Engineering  Science  Series 


ENGINES   AND   BOILERS 


ENGINEERING  SCIENCE  SERIES 

EDITED  BY 
DUGALD  C.  JACKSON,  C.E. 

PROFESSOR  OF  ELECTRICAL  ENGINEERING 

MASSACHUSETTS  INSTITUTE  OF  TECHNOLOGY 

FELLOW  AND  PAST  PRESIDENT  A.I.E.E. 

EARLE  R.  HEDRICK,  Ph.D. 

PROFESSOR  OF  MATHEMATICS,  UNIVERSITY  OF  MISSOURI 
MEMBER  A.S.M.E. 


ENGINES  AND  BOILERS 


BY 

THOMAS  T.    EYRE 

DEAN,  COLLEGE  OF  ENGINEERING 

STATE    UNIVERSITY   OF   NEW   MEXICO, 

FORMERLY    ASSISTANT    PROFESSOR   OF 

MECHANICAL  ENGINEERING,  PURDUE  UNIVERSITY 


J?eto  gorfe 

THE  MACMILLAN  COMPANY 
1922 

All  rights  reserved 


PRINTED   IN   THE   UNITED   STATES   OF   AMERICA 


COPYRIGHT,  1922, 
BY  THE  MACMILLAN  COMPANY 


Set  up  and  electrotyped.    Published  August,  1922. 

.  -  FORESTRY 


Press  of  J.  J.  Little  &  Ives  Co. 
New  York 


/  J  ^ 
,£  1 

Yo  ^es^vv 


PREFACE 

This  text  book  on  Engines  and  Boilers  is  intended  for  use  in 
engineering  schools  which  offer  an  elementary  course  in  Heat 
Engines.  No  attempt  has  been  made  to  cover  the  more  advanced 
work  in  Thermodynamics,  or  to  give  an  exhaustive  treatment  of 
the  subject  of  Heat  Power. 

This  work  is  the  result  of  the  author's  experience  during  the 
several  years  that  he  taught  classes  in  Engines  and  Boilers  and  in 
allied  subjects  at  Purdue  University.  Much  of  the  material  was 
given  to  the  students  first  in  lectures,  and  later  in  the  form  of 
mimeographed  notes.  It  is  now  presented  in  book  form  with  the 
hope  that  it  may  be  of  value  in  other  engineering  schools. 

At  the  end  of  the  book  a  list  of  representative  problems  is  given. 
It  has  been  the  author's  experience  that  the  student  obtains  a 
better  understanding  of  the  subject  if  he  is  required  to  work 
problems  related  to  the  matter  in  the  text. 

The  author  wishes  to  thank  Professor  C.  H.  Lawrence  for  valu- 
able suggestions  made  in  regard  to  the  form  of  presentation  of  some 
of  the  work. 

THOMAS  T.  EYRE. 

University  of  New  Mexico. 


M512399 


CONTENTS 

CHAPTER  I 

PAGE 

PRESSURE,  TEMPERATURE  AND  HEAT  UNITS     ...  1 

CHAPTER   II 

FUEL      .  ...  4 

Anthracite  Coal  —  Bituminous  Coal  —  Lignite  and  Peat  — 
Natural  Gas  —  Oil  —  Coal  Fields  of  the  U.  S.  —  Coal  Storage  - 
Determination  of  the  Heating  Value  of  Fuel  —  Combustion  — 
Composition  of  Flue  Gas  —  Flue  Gas  Analysis  —  Heat  Lost  in  Flue 
Gas  —  Value  of  CO2  for  Best  Efficiency  —  CO2  Recorders. 

CHAPTER  III 

STEAM    .....  ....       16 

Use  of  Steam  Tables  —  Throttling  Calorimeter. 

CHAPTER  IV 

BOILERS          ...........       23 

Requirements  —  Rated  Horsepower  —  Heating  Surface  —  Rules 
for  Finding  the  Heating  Surface  —  Superheating  Surface  —  Size 
of  Boiler  Tubes  —  The  B.  &  W.  Boiler  —  The  Sterling  Boiler  —  The 
Wickes  Boiler  — The  Return  Tubular  Boiler  —  The  Internally- 
fired  Return  Tubular  Boiler  —  The  Scotch  Marine  Boiler  —  The 
Vertical  Fire  Tube  Boiler  —  The  Locomotive  Boiler  —  Superheaters 

—  Horsepower  of  boilers  —  Factor  of  Evaporation  —  Efficiency  of 
Boilers  —  A.  S.  M.  E.  Boiler  Test  Code. 

CHAPTER  V 

BOILER  ACCESSORIES  AND  AUXILIARIES     ......       45 

Grates  —  The  Plain  Grate  —  The  Rocking  Grate  —  Mechanical 
Stokers  —  The  Chain  Grate  —  The  Roney  Stoker  —  The  Under- 
feed Furnace  —  Smoke  Prevention  —  Settings  —  Draft  —  Dam- 
pers —  Safety  Devices  —  The  Pressure  Gage  —  The  Safety-valve 

—  Safety-valve  Capacity  —  Napier's  Formula  —  Safety-valve  For- 
mula —  The  Water  Glass  or  Gage  Glass  —  High-water  and  Low- 
water  Alarm  —  The  Fusible  Plug  —  Boiler  Feedwater  Treatment 

—  Scale  Prevention  and  Removal  —  Oil  Separators  —  Boiler  Feed 
Pumps  —  The  Injector  —  Boiler  Feed  by  Returning  Trap  —  The 
Steam  Line  —  The  Steam  Trap  —  Expansion  Joints  —  Steam 

vii 


Vlll  CONTENTS 

PAGE 

Separators  —  Steam-pipe  Covering  —  Feedwater  Heaters  —  Econ- 
omizers —  Condensers  —  The  Surface  Condenser  —  The  Jet  Con- 
denser —  Cooling  of  Circulating  Water. 

CHAPTER  VI 

THE  STEAM  ENGINE 76 

History  —  The  Plain  Slide-valve  Engine  —  Parts  of  the  Steam 
Engine  —  Piston  Displacement  —  Clearance  —  Steam  Back  of 
Piston  During  Stroke  —  The  Indicator  and  its  Purposes  —  Events 
of  Stroke  — •  Location  of  Events  on  Diagram  —  Equation  of  Expan- 
sion and  Compression  Curves  —  Hypothetical  Indicator  Diagram  — 
Determination  of  Clearance  from  Card  —  Determination  of  the 
Mean  Effective  Pressure  —  Indicated  Horsepower  —  Brake  Horse- 
power —  Mechanical  Efficiency  —  Thermal  Efficiency  —  Cylinder 
Condensation  —  Steam  accounted  for  by  the  Indicator  Diagram  — 
Valve  Setting  from  the  Indicator  Diagram. 

CHAPTER  VII 

COMMON  TYPES  OF  STEAM  ENGINES 100 

Slide-valve  Engine  —  The  Corliss  Engine  —  The  Four-valve 
Engine  —  The  Compound  Engine  —  The  Tandem-compound  — 
The  Cross-compound  —  Cylinder  Ratio  —  The  Combined  Indi- 
cator Diagram  —  The  Diagram  Factor  —  Ratio  of  Expansion  — 
The  Unaflow  Engine. 

CHAPTER   VIII 

VALVES 112 

The  D  Slide-valve  —  Relative  Motion  of  Crank  and  Piston  — 
Valve  Diagrams  —  The  Valve  Ellipse  —  The  Bilgram  Diagram  — 
The  Zeuner  Diagram  —  Types  of  Slide-valves  —  Valve  with  Pres- 
sure Plate  — The  Piston  Valve  —  Double-ported  Valves  —  The  ' 
Gridiron  Valve  —  The  Riding  Cut-off  Valve  —  Effect  of  Rocker  Arm 
on  the  Location  of  Eccentric  —  Oscillating  Valves  —  Poppet  Valves 
—  Reversing  —  The  Stephenson  Link  —  The  Walschaert  Valve 
Gear  —  The  Joy  Valve  Gear—  Setting  the  Slide-valve. 

CHAPTER  IX 

GOVERNORS    ...........     137 

General  —  Classification  of  Governors  —  The  Gravity-balanced 
Spindle  Governor  —  The  Spring-balanced  Governor  —  Governing 
by  Changing  Position  of  Eccentric  —  Governing  by  Changing  Angle 
of  Advance  —  Governing  by  Changing  Both  Angle  of  Advance  and 
Valve  Travel  —  Centrifugal  and  Inertia  Governors. 

CHAPTER  X 

STEAM  TURBINES 150 

General  —  History  —  Fundamental  Principles  —  Available  En- 
ergy in  Steam  —  Velocity  Due  to  Expansion  —  Impulse  and  Reac- 


CONTENTS  lx 

PAGE 

tion  —  Bucket  Shapes  —  Single-stage  Turbine  —  Staging  —  Multi- 
stage Impulse,  Velocity-staging  —  Multi-stage  Impulse,  Pressure- 
staging  —  Multi-stage  Impulse,  Combination  Pressure-staging  and 
Velocity-staging  —  Multi-stage  Reaction  —  Change  of  Area  of 
Steam  Passage  —  Leakage  —  Loss  due  to  Running  Partial  Capacity 
—  Summary  of  Losses  in  the  Steam  Turbine  —  Common  Commer- 
cial Types  —  DeLaval  Single-stage  —  Multi-pressure-stage  Im- 
pulse Turbine  —  Curtis  Turbine  —  Parsons  Turbine  —  Westing- 
house  Turbine  —  Other  Types  —  Low-pressure  Turbine  —  Mixed- 
pressure  Turbine  —  Use  of  Superheated  Steam  —  The  Marine 
Turbine. 

CHAPTER  XI 

GAS  ENGINES          ..........     193 

General  —  Historical  —  Cycles  of  Operation  —  The  Four-stroke 
Cycle  —  The  Two-stroke  Cycle  —  Classification  from  Fuel  Used  — 
Efficiency  —  Fuels  —  The  Gasoline  Carburetor  —  The  Gas  Pro- 
ducer—  Cooling  of  Cylinders  —  Ignition  —  Valves  —  Governing 
—  Determination  of  Horsepower  —  Multi-cylinder  Engines. 

PROBLEMS  213 


ENGINES  AND  BOILERS 

CHAPTER  I 
PRESSURE,    TEMPERATURE  AND    HEAT  UNITS 

1.  Pressure  Units.  —  In  steam  engineering,  pressure  is  meas- 
ured in  the  following  units: 

(1)  In  pounds  per  square  inch. 

(2)  In  inches  of  mercury. 

(3)  In  inches  of  water. 

In  this  country  boiler  pressures  are  ordinarily  measured  in 
pounds  per  square  inch  above  atmospheric  pressure.  Condenser 
pressures  are  commonly  measured  from  atmospheric  pressure  in 
inches  of  mercury,  i.e.  the  difference  between  the  pressure  in 
the  condenser  and  atmospheric  pressure  is  read  on  a  mercury 
column.  Draft  pressures  are  usually  measured  in  inches  of  water. 
Pressure  gages  and  vacuum  gages  are  so  constructed  that  they 
read  zero  at  atmospheric  pressure.  The  atmosphere  exerts  a  vari- 
able pressure,  which  is  about  14.5  pounds  per  square  inch  at  ordi- 
nary altitudes  and  under  ordinary  conditions.  At  sea-level,  the 
standard  is  taken  as  14.7  pounds  per  square  inch,  which  is  equiv- 
alent to  29.92  inches  of  mercury. 

As  the  atmospheric  pressure  is  slightly  variable,  it  is  necessary 
in  accurate  work  that  pressures  be  reduced  to  an  absolute  basis. 
Since  the  zero  reading  of  a  boiler  pressure  gage  means  atmospheric 
pressure,  the  absolute  pressure  will  be  the  sum  of  the  gage  pressure 
and  atmospheric  pressure. 

A  partial  vacuum  usually  exists  in  a  condenser,  since  the  abso- 
lute condenser  pressure  is  usually  less  than  atmospheric  pressure. 
Since  the  vacuum  gage  as  well  as  the  boiler  pressure  gage  reads 
zero  at  atmospheric  pressure,  the  absolute  pressure  in  the  con- 
denser is  the  difference  between  the  atmospheric  pressure  and  the 
vacuum-gage  pressure. 

Figure  1  shows  diagrammatically  the  relation  between  these 
pressures.  In  this  figure,  a  is  the  boiler-gage  pressure,  b  the 

1 


2  ENGINES   AND   BOILERS 

atmospheric  pressure,  and  c  the  absolute  boiler  pressure.  Like- 
wise, d  represents  the  vacuum-gage  pressure,  which  is  measured 
downward  from  the  atmospheric  pressure;  and  e,  the  difference 
between  6  and  d,  is  the  absolute  condenser  pressure. 


I 

1 

a        1 

:    r 

t      i 

6 

d 

\ 

l_L_  Condenser  pressure 

V             f 

£—  Z^/»  pressure 

FIG.  1 

Barometers  used  in  engineering  work  in  this  country  are  usu- 
ally graduated  in  inches.  The  mercury  barometer  is  used  be- 
cause mercury  is  the  most  convenient  fluid  for  this  use.  A  cubic 
inch  of  mercury  weighs  very  nearly  0.49  of  a  pound.  Hence  the 
pressure  in  pounds  per  square  inch  is  the  barometric  reading  in 
inches  multiplied  by  0.49  Vacuum  gages  also  are  commonly 
graduated  to  read  in  inches  of  mercury.  Therefore  the  absolute 
condenser  pressure  in  pounds  per  square  inch  is  0.49  times  the 
difference  between  the  barometer  and  the  vacuum-gage  readings. 

EXAMPLE  1.  Find  the  absolute  boiler  pressure  when  the  pressure  gage 
reads  110  pounds  and  the  barometer  reads  29.4  inches. 

SOLUTION.  When  the  barometer  reads  29.4  inches,  the  atmospheric  pres- 
sure is  29.4  X0.49  =  14.4  pounds  per  square  inch.  The  absolute  boiler  pressure 
is  then  14.4  +  110  =  124.4  pounds  per  square  inch. 

EXAMPLE  2.  Find  the  absolute  pressure  in  a  condenser  when  the  barometer 
reads  29.8  inches  and  the  vacuum  gage  reads  27.3  inches. 

SOLUTION.  The  absolute  condenser  pressure  is  the  difference  between  the 
barometric  and  the  vacuum-gage  pressures,  which  in  inches  of  mercury  is 
29.8-27.3  =  2.5.  This  reduced  to  pounds  per  square  inch  is  2.5.  X49  =  1.22. 

2.  Temperature  Units.  —  Ordinary  temperatures  are  measured 
by  means  of  the  mercury  thermometer.  For  higher  tempera- 
tures, such  as  those  that  occur  in  furnaces,  special  thermometric 
devices  called  pyrometers  are  used. 

Three  thermometer  scales  are  in  use,  the  Fahrenheit,  the  Cen- 
tigrade, and  the  Reaumur.  In  the  Fahrenheit  scale,  the  differ- 
ence between  the  temperatures  of  melting  ice  and  boiling  water 
at  sea  level  is  divided  into  180  divisions  or  degrees;  the  freezing 
point  is  32°  and  the  boiling  point  212°.  This  makes  the  zero 


PRESSURE,    TEMPERATURE,    AND    HEAT   UNITS 


Fahrenheit    Centigrade        Reaumur 


point  come  32°  below  the  freezing  point  of  water.  In  the  Centi- 
grade scale,  the  freezing  point  is  0°  and  the  boiling  point  is  100°. 
In  the  Reaumur  scale,  the  freezing  point  is  0°  and  the  boiling 
point  is  80°.  Figure  2  shows  graphically  the  relation  between  these 
scales.  It  is  readily  seen 
how  the  reading  on  one 
scale  may  be  reduced  to 
that  of  either  of  the 
others.  It  is  serviceable 

-„•-!.  - 

|  67/.S- 


1 


80'  i  Boili/tf  temperature 


femperofare 


\\        '' 

«-J \  A  Absolute  zero 

FIG.  2 


to  remember  that  a 
difference  of  temperature 
equal  to  5°  on  the  Cen- 
tigrade scale  is  equal  to 
9°  on  the  Fahrenheit 
scale. 

Experiment  shows 
that  a  perfect  gas  under 
constant  pressure  at  32°  Fahrenheit  expands  1/491.5  part  of  its 
own  volume  for  each  degree  (F.)  that  its  temperature  is  in- 
creased. Because  of  this  we  call  the  point  491. 5°  below  the  freez- 
ing point,  or  —459.5°  on  the  Fahrenheit  scale,  the  absolute  zero. 
This  corresponds  to  a  Centigrade  temperature  of  —  273°. 

3.  Heat  Units.  —  In  engineering  work  in  this   country,  it  is 
customary  to  use  English  units.     Our  common  heat  unit  is  the 
British  thermal  unit  (B.t.u.),  which  is  practically  defined  as  the 
amount    of   heat   necessary    to    raise    the    temperature    of   one 
pound  of  pure  water  from  62°  to  63°  Fahrenheit.     In  the  metric 
system  the  engineers'  heat  unit  is  the  large  calorie,  which  is  the 
amount  of  heat  necessary  to  raise  the  temperature  of  a  kilogram 
of  water  from  15°  to  16°  Centigrade.     As  one  kilogram  =2.2046 
pounds,  and  as  1°  Centigrade  =1.8  degrees  Fahrenheit, 

one  calorie  =3. 968  B.t.u.,  or  one  B.t.u.  =0.252  calorie. 

4.  Mechanical  Equivalent  of  Heat.  —  Heat  experiments  have 
been  made  to  determine  the  relation  between  heat-energy  and 
mechanical  work.     The  latest  and  most  refined  experiments  show 
that  one  B.t.u.  as  above  defined  is  substantially  equivalent  to  778 
foot-pounds  of  work.     This  relation  is  called  the  mechanical  equiva- 
lent of  heat.1 

*For  definition  of  the  "mean  B.  t.  u."  and  corresponding  Mechanical    Equivalent  of  Heat 
see  A.  S.  M.  E.  POWER  TEST  CODE,  Edition  of  1915,  p.  28. 


CHAPTER  II 
FUEL 

5.  Introduction.  —  The  source  of  the  world's  supply  of  energy 
is  the  sun.  In  the  use  of  water  power  we  are  drawing,  in  point 
of  time,  almost  directly  from  the  sun.  On  the  other  hand,  in 
the  use  of  such  fuels  as  coal,  gas,  and  oil,  we  make  use  of  a  store 
of  energy  that  has  been  accumulating  for  ages. 

While  the  world's  supply  of  coal,  oil,  and  gas  is  limited,  we  have 
used  but  a  very  small  part  of  the  known  deposits.  The  past 
century  has  seen  a  marvelous  change  in  our  manner  of  living  and 
in  our  ways  of  thinking.  Our  vast  commercial  system  with  its 
perplexing  problems  has  arisen  during  the  past  few  generations. 
One  of  the  chief  causes  of  this  great  change  is  our  ability  to  util- 
ize the  vast  stores  of  energy  to  be  found  in  nature.  The  steam- 
engine  has  been  the  chief  means  by  which  the  energy  stored  in 
our  coal  deposits  has  been  tapped  and  forced  to  do  the  work  of 
man.  What  will  become  of  this  modern  civilization  of  ours  when 
the  fuel  supply  upon  which  it  is  based  is  exhausted,  is  an  inter- 
esting problem  of  the  future.  Already  conservationists  are  call- 
ing to  us  to  stop  the  great  waste  of  our  natural  resources. 

Of  all  our  fuels,  coal  is  the  most  important.  Coal  is  the  re- 
mains of  vegetable  matter  deposited  in  remote  geological  ages. 
It  is  well  known  that  wood  rots  but  little  when  kept  under  water. 
If  the  water  be  fresh,  -the  wood  is  not  eaten  by  the  teredo  worm 
or  other  forms  of  aquatic  life,  and  will  be  kept  in  a  fair  state  of 
preservation  for  thousands  of  years.  If  tree  trunks  and  other 
vegetable  matter  fall  into  a  fresh-water  swamp  and  are  sub- 
merged before  they  rot,  and  if  this  continues  for  many  centuries, 
there  will  be  a  great  accumulation  of  it.  Oar  coal  deposits  are 
the  result  of  such  an  accumulation  of  vegetable  matter.  Under 
tropical  conditions  accompanied  by  a  large  supply  of  carbon 
dioxide  in  the  atmosphere  the  growth  was  very  rapid  and  a  deep 
bed  would  collect  in  a  comparatively  short  time. 

Geologists  tell  us  that  in  the  past,  parts  of  the  surface  of  the 
earth  have  gradually  risen  while  others  have  fallen.  At  a  remote 
time,  the  tops  of  our  highest  mountains  may  have  been  the  bot- 
tom of  the  sea.  Suppose  a  former  swamp  with  its  accumulated 
vegetable  matter  is  now  sunk,  and  that  great  quantities  of  silt  or 

4 


FUEL  5 

other  material  are  deposited  upon  it.  The  weight  of  the  material 
above  will  compress  the  vegetable  matter  into  a  compact  and 
dense  mass.  It  is  also  possible  that  it  will  be  subject  to  a  high 
temperature,  which  will  change  its  chemical  composition.  Vary- 
ing conditions  of  pressure  and  heat  are  thought  to  be  responsible 
largely  for  the  differences  between  the  various  kinds  of  coal. 

The  principal  constituents  of  coal  are  carbon,  hydrogen,  oxygen, 
nitrogen,  sulphur,  and  refractory  earths  called  ash.  The  wood- 
fiber  of  the  original  vegetable  matter  was  composed  chiefly  of 
hydrocarbons.  While  under  the  influence  of  great  pressure,  it 
has  at  some  period  of  its  history  been  subjected  to  considerable 
heat  and  therefore  undergone  a  process  of  destructive  distiilization. 
This  has  driven  off  much  of  the  volatile  matter  of  the  original 
vegetable  material  and  left  a  considerable  portion  of  uncombined 
carbon,  which  is  called  fixed  carbon.  The  remainder  of  the  car- 
bon exists  in  combination  with  hydrogen.  These  carbon  and 
hydrogen  compounds  are  called  hydrocarbons.  They  are  easily 
volatilized,  and  so  comprise  a  part  of  the  volatile  matter  of  the 
coal. 

Oxygen  and  hydrogen  are  always  present  in  coal  in  the  form 
of  water.  This  water  is  volatilized,  of  course,  when  the  coal  is 
burned.  Since  heat  is  required  to  evaporate  and  to  superheat  it, 
water  is  a  detriment  if  present  in  too  large  a  quantity.  A  small 
amount  of  water,  however,  seems  to  aid  in  the  combustion  of 
some  coals. 

All  coal  then  contains  fixed  carbon,  volatile  matter  (hydrocar- 
bons and  water),  and  ash;  and  it  may  contain  other  substances 
(e.g.,  sulphur).  The  combustible  is  the  fixed  carbon,  the  hydro- 
carbons, and  part  of  the  sulphur  that  may  be  present.  Excessive 
sulphur  is  undesirable  because  it  is  harmful  to  the  metal  of  the 
boiler  and  the  stack  if  moisture  is  present,  since  it  may  form 
sulphurous  or  sulphuric  acid;  it  combines  with  the  ash  to  form 
a  fusible  slag  or  clinker,  which  is  commonly  objectionable;  and 
it  makes  the  fuel  more  liable  to  spontaneous  combustion  when 
stored  in  deep  piles. 

Coals  that  have  been  subjected  to  the  greatest  pressure  and 
heat  are  composed  mostly  of  fixed  carbon,  and  contain  only  a 
small  amount  of  volatile  hydrocarbons.  Such  coals  are  called 
anthracite.  Coals  containing  larger  quantities  of  volatile  hydro- 
carbons are  called  bituminous.  Since  there  is  no  definite  divid- 


6  ENGINES   AND    BOILERS 

ing  line  between  these  two  classes,  but  the  two  seemingly  overlap, 
the  terms  semi-anthracite  and  semi-bituminous  are  commonly  used 
to  designate  coals  to  which  are  ascribed  certain  properties  of 
each  class. 

6.  Anthracite  Coal.  —  Anthracite  coal  contains  but  a  small 
amount  of  combustible  volatile  matter.     While  it  is  considered 
better  for  some  uses,  its  heating  value  is  less  than  that  of  good 
grades  of  bituminous  coal.     It  burns  slowly,  with  but  a  small 
flame  and  practically  no  smoke.    Due  to  its  slow  burning  quali- 
ties, a  relatively  large  grate  area  is  needed  on  which  to  burn  it. 
The  supply  of  this  coal  that  is  easily  obtained  has  been  diminish- 
ing in  this  country,  and  its  demand  for  domestic  use  has  greatly 
increased  during  the  past  few  years.     This  has  led  to  a  rapid  in- 
crease in  price  and  a  great  diminution  of  its  use  for  power  purposes. 

Anthracite  is  considered  much  superior  to  bituminous  coal  for 
the  production  of  producer  gas.  This  is  due  to  the  fact  that  it  is 
so  free  from  the  hydrocarbons  that  produce  tars.  The  formation 
of  tar  has  been  the  great  objection  to  the  use  of  bituminous  coals 
in  the  producer  plant. 

Average  anthracite  coal  contains  about  85%  fixed  carbon,  4% 
volatile  matter,  9%  ash,  and  2%  water.  The  heat  value  averages 
about  13000  B.t.u.  per  pound. 

7.  Bituminous  Coal.  —  Most  of  the  coal  used  for  power  pur- 
poses is  bituminous.     This  coal  contains  a  larger  percentage  of 
volatile  combustible  matter.     It  burns  at  a  lower  temperature 
than  does  anthracite,  and  with  a  much  longer  flame.    The  length 
of  the  flame  varies  with  the  composition,  some  kinds  being  called 
long-flaming  and  others  short-flaming.     The   ordinary  furnace  is 
not  usually  so  constructed  as  to  give  the  volatile  hydrocarbons 
a  chance  to  be  completely  burned.    This  results  in  the  formation 
of  smoke.     Engineers  have  spent  much  time  and  study  on  the 
prevention  of  smoke.     Upon  yielding  up  their  volatile  matter 
some  coals  fuse  and  form  a  solid  mass  or  cake  of  the  nature  of 
coke.    These  are  called  caking  coals.    This  action  hinders  the  draft. 
If  rapid  combustion  is  desired,  the  mass  must  be  broken  up. 

The  composition  of  bituminous  coal  varies  greatly,  but  the 
average  of  the  better  grades  may  be  taken  as  65%  fixed  carbon, 
28%  volatile  combustible,  5%  ash,  and  2%  moisture,  with  a  heat 
value  of  14000  B.t.u.  per  pound. 


FUEL  7 

8.  Lignite  and  Peat.  —  Lignite  or  brown  coal  is  high  in  vola- 
tile combustible  and  also  contains  much  moisture.     The  evapora- 
tion of  this  moisture  after  mining  causes  the  lignite  to  crumble  or 
slack.     It  is  usually  inferior  to  anthracite  and  bituminous  coals, 
but  it  is  used  where  it  is  easily  obtained  and  where  better  coal  is 
expensive.     Abroad,  lignite  is  often  formed  into  briquettes. 

Peat  is  the  partly  decayed  remains  of  vegetation  that  accumu- 
lates in  bogs.  While  it  is  an  inferior  fuel,  it  is  used  to  a  consid- 
erable extent  abroad.  It  is  sometimes  pressed  into  briquettes. 

9.  Natural  Gas.  —  Natural  gas  is  used  to  some  extent  for  power 
purposes  in  sections  of  the  country  within  reach  of  the  gas  fields. 
It  contains  about  90%  marsh  gas  (CH4)  and  has  a  heat  value  of 
nearly  1000  B.t.u.  per  cubic  foot. 

10.  Oil.  —  Crude  petroleum  and  fuel  oil  are  used  to  a  consid- 
erable extent  in  parts  of  this  country.     The  petroleum  produced 
in  the  eastern  and  middle  states  is  of  a  paraffin  base,  while  that 
from  Texas  and  California  is  of  an  asphalt  base.     Gasoline  and 
other  light  oils  are  distilled  from  petroleum,  and  the  residue  is 
sold  as  fuel  oil.     Petroleum  is  composed  principally  of  hydrogen 
and  carbon  in  the  form  of  hydrocarbons  and  has  a  heat  value  of 
about  20000  B.t.u.  per  pound. 

11.  Coal  Fields  of  the  United  States.  —  Our  principal  deposits 
of  anthracite  coal  are  in  eastern  Pennsylvania.     The  deposits  are 
not  of  large  extent  and  the  best  are  rapidly  becoming  exhausted. 
It  is  claimed  that  there  is  anthracite  in  Alaska.     Only  a  small 
proportion  of  the  anthracite  mined  is  being  used  for  power  pur- 
poses, the  rest  going  for  domestic  heating  and  like  purposes. 

The  best  of  our  bituminous  and  semi-bituminous  coal  is  taken 
from  the  field  that  includes  western  Pennsylvania,  eastern  Ohio, 
a  large  part  of  West  Virginia,  and  eastern  Kentucky.  Southwest- 
ern Indiana  and  most  of  the  state  of  Illinois  are  underlaid  with  coal 
of  a  fair  quality.  There  is  also  a  field  running  north  from  Okla- 
homa through  eastern  Kansas  and  western  Missouri  into  Iowa. 
The  coal  from  the  latter  is  generally  of  poor  quality,  and  is  used 
only  locally.  Immense  coal  fields  exist  in  southern  Utah  and 
Colorado,  and  in  New  Mexico  and  Arizona.  No  complete  survey 
of  these  fields  has  been  made  as  yet,  and  they  have  not  been  de- 
veloped up  to  the  present.  There  is  also  a  coal  and  lignite  field 
in  eastern  Montana  and  western  North  Dakota. 


8  ENGINES   AND    BOILERS 

The  peat  beds  of  the  country  are  principally  in  Minnesota, 
Wisconsin,  Michigan,  New  York,  and  the  New  England  states. 

Oil  is  produced  in  the  territory  occupied  by  the  eastern  coal 
fields,  in  Kansas,  Oklahoma,  Texas,  California,  and,  to  a  smaller 
extent,  elsewhere. 

12.  Coal  Storage.  —  In  the  operation  of  most  steam-power 
plants,  it  is  essential  that  a  constant  supply  of  coal  be  available. 
Owing  to  unsettled  labor  conditions  at  the  mines  and  to  uncer- 
tain transportation  facilities,  it  is  necessary  that  there  be  some 
storage  capacity.    With  anthracite  coal  this  is  a  simple  problem, 
but  it  is  not  so  with  certain  grades  of  bituminous  coal.     Upon 
exposure  of  bituminous  coal  to  the  air,  there  is  a  considerable 
oxidation  of  the  hydrocarbons  with  attendant  heat  production. 
If  the  coal  pile  is  large,  this  heat  may  start  a  fire  which  is  costly 
and  hard  to  extinguish.    The  origin  of  such  a  fire  is  called  spon- 
taneous combustion.    Even  if  fire  does  not  start,  there  is  a  loss  of 
heat-value  up  to  as  high  as  10%  in  some  grades  of  coal.    If  the 
moisture  content  is  large,  the  weathering  is  accompanied  by  a 
crumbling  or  slacking.    Storage  piles  are  often  ventilated  in  order 
to  keep  them  cool.     In  some  large  plants,  the  storage  is  so  ar- 
ranged that  it  may  be  submerged  in  water.     This  obviates  the 
fire  risk,  and  reduces  the  other  losses  to  a  minimum. 

13.  Determination  of  Heating  Values  of  Fuel.  —  In   plants 
where  large  amounts  of  fuel  are  used,  it  is  quite  common  to  buy 
coal  on  the  basis  of  its  heating  value.    In  accurate  tests  of  power 
plants  it  is  also  necessary  to  know  the  heating  value  of  the  coal 
used.    Care  must  be  exercised  in  order  to  get  a  fair  sample  of  the 
fuel.     The  heating  value  may  be  determined  in  two  ways,  by 
combustion  in  a  calorimeter,  or  by  chemical  analysis. 

In  the  calorimeter  method  a  sample  of  the  fuel  is  placed  in  a 
steel  bomb  along  with  compressed  oxygen.  The  bomb  is  placed 
in  a  calorimeter  containing  water,  and  the  fuel  is  ignited  by  means 
of  an  electrically  heated  wire.  Upon  the  firing  of  the  fuel,  heat 
is  given  to  the  water.  It  is  possible  to  determine  from  the  rise 
of  the  temperature  of  the  water  the  amount  of  heat  generated. 
The  value  thus  determined  is  known  as  the  higher  calorific  value. 
It  must  be  remembered  that  it  is  seldom  possible  actually  to  get 
this  amount  of  heat  by  burning  in  a  furnace.  This  is  due  to  the 
fact  that  the  hydrogen  in  the  fuel  combines  with  the  oxygen  of 


FUEL  9 

the  air  to  form  water,  which  ordinarily  passes  from  the  furnace 
in  the  form  of  steam,  carrying  with  it  the  heat  of  vaporization 
of  the  steam.  The  lower  heat-value  does  not  contain  this  heat 
of  vaporization.  The  higher  value  is  the  accepted  standard. 

There  are  two  methods  of  chemical  analysis,  the  ultimate,  and 
the  proximate.  The  ultimate  analysis  may  be  made  on  the  basis 
either  of  moist  or  dry  fuel.  The  latter  is  commonly  accepted. 
If  the  analysis  is  made  on  the  basis  of  moist  fuel,  it  may  be  con- 
verted to  the  dry  basis  by  dividing  the  percentage  of  the  various 
constituents  by  one  minus  the  percentage  of  the  moisture.  The 
ultimate  analysis  gives  the  percentage  by  weight  of  carbon,  hydro- 
gen, oxygen,  nitrogen,  sulphur,  and  ash.  Knowing  the  composi- 
tion of  the  fuel,  the  heat-value  may  be  determined  by  a  formula. 
An  accepted  formula  is  a  modification  of  that  of  DULONG  : 

B.t.u.  per  pound  of  dry  fuel  =  14600  C+62000  (H-O/8)-f  4000S, 

in  which  C,  H,  O,  and  S  represent  the  proportionate  parts  by 
weight  of  carbon,  hydrogen,  oxygen,  and  sulphur.  The  heat- 
value  of  pure  sulphur  is  4000  B.t.u.  per  pound,  but  the  sulphur 
in  coal  is  mostly  in  a  form  that  is  noncombustible.  The  results 
of  the  ultimate  analysis  agree  closely  with  those  obtained  from 
the  calorimeter.  The  proximate  analysis  gives  the  proportion  of 
fixed  carbon,  volatile  combustible,  moisture,  and  ash.  Since  the 
heating  value  of  the  volatile  combustible  is  not  determined,  the 
results  are  not  as  reliable  as  those  of  the  previous  method.  It  is 
very  often  used,  however,  because  it  is  easily  made  and  affords  a 
rough  comparison  of  various  fuels. 

In  making  the  proximate  analysis,  a  weighed  sample  of  coal 
is  placed  in  a  crucible  of  known  weight  and  is  kept  at  a  tem- 
perature a  few  degrees  above  the  boiling  point  of  water  for  an 
hour.  From  the  loss  of  weight,  the  moisture  in  the  coal  is  cal- 
culated. The  sample  is  next  heated  in  a  hot  flame  a  few  minutes 
with  the  lid  on  the  crucible.  This  drives  off  the  volatile  matter, 
and  the  difference  in  the  new  and  previous  weight  is  the  amount 
driven  off.  Now  the  sample  is  heated  for  two  hours  with  the  lid 
off,  all  fixed  carbon  is  burned  out,  and  the  weight  is  determined 
by  difference  as  before.  The  weight  left  is  that  of  the  ash.  After 
having  found  the  percent  by  weight  of  the  moisture,  volatile 
matter,  fixed  carbon,  and  ash,  the  approximate  calorific  value 
may  be  found  by  means  of  the  chart,  shown  in  Fig.  3,  which 


10 


ENGINES   AND   BOILERS 


has  been  constructed  from  the  values  determined  in  a  large  num- 
ber of  accurate  analyses.  On  this  chart  the  heat-value  in  B.t.u. 
per  pound  of  combustible  is  plotted  against  the  percent  of  fixed 
carbon  in  the  total  combustible.  It  is  assumed  that  the  volatile 
matter  and  the  fixed  carbon  constitute  the  combustible,  the  ash 
and  the  moisture  being  noncombustible.  The  method  of  getting 
the  heating  value  may  be  shown  by  examples. 

EXAMPLE  1.  Determine  the  calorific  value  in  B.t.u.  per  pound  of  dry 
coal  having  the  following  ultimate  analysis:  carbon  =  75.46%,  hydrogen  = 
3.34%,  oxygen  =  2.70%,  nitrogen  =  .53%,  sulphur  =  2.54%,  and  ash  =  15.43%. 

SOLUTION.  The  B.t.u.  per  pound  =  14600  X  .7546 +62000  (.0334  -  .0270/8) 
+400t<  .0254  =  12880  B.t.u.  The  oxygen  calorimeter  gave  a  value  of  13000 
B.t.u.  per  pound  for  this  same  sample. 

Chart  for  determining  Heat  Value  of  Combustible  with  Different  Percentages 
of  Fixed  Carbon  from  Proximate  Analysis 


^ 

^ 

•^ 

^ 

"X 

S 

x 

s 

\ 

3+O0 

s 

\ 

s 

\! 

S 

\ 

| 

/ 

/ 

^ 

^        /  <•/  a/t 

/ 

/3/QQ 

\ 

\^        '*><*" 

/ 

s 

/ 

/ 

\ 

^       14000 

s 

J 

^ 

^         t+fVQ 

f 

\ 

/ 

V              *  r 

1 

1 

* 

1 

I 

tot 

^      -14,00 

1 

14400 

JO  SS  £0  6S  70  7S  &0 

Percent  of  ftxcd  C  arbor?  m  Tbta/ 
FIG.  3 

EXAMPLE  2.  Determine  the  calorific  value  of  coal  which  has  the  follow- 
ing proximate  analysis:  moisture  =4. 7%,  volatile  matter  =  24.6%,  fixed 
carbon =62.5%,  and  ash  =  8.2%. 

SOLUTION.  The  combustible  being  composed  of  the  volatile  matter  and 
fixed  carbon  comprises  24.6+62.5  =  87.1%  of  the  weight  of  the  coal.  Of 
this  combustible  the  fixed  carbon  is  62.5/87.1  =  71.7%.  From  the  chart 
it  is  seen  that  for  71.7%  the  B.t.u.  per  pound  of  combustible  is  15550  B.t.u. 
As  the  combustible  comprises  only  87.1%  of  the  total  weight  the  heating 
value  will  be  .871X15550  =  13550  B.t.u.  per  pound  of  wet  coal,  or  if  reduced 
to  the  dry  basis,  13550/(1.00-. 047)  =  14200  B.t.u.  per  pound. 


FUEL  11 

14.  Combustion.  —  By  combustion  is  meant  the  rapid  chem- 
ical combination  of  oxygen  with  the  carbon,  hydrogen,  and  sul- 
phur in  fuel.  Combustion  is  complete  when  the  maximum  amount 
of  oxygen  is  used  in  the  combination.  One  atom  of  carbon  will 
combine  with  one  atom  of  oxygen  to  form  carbon  monoxide  (CO). 
This  is  not  complete  combustion,  however,  for  one  atom  of  car- 
bon will  combine  with  two  atoms  of  oxygen,  forming  carbon  di- 
oxide (CO2),  if  sufficient  oxygen  is  present. 

Since  the  atomic  weight  of  oxygen  is  16,  and  that  of  carbon  is 
12,  it  takes  32  pounds  of  oxygen  to  12  pounds  of  carbon  to  form 
carbon  dioxide,  i.e.,  one  pound  of  carbon  requires  for  its  complete 
oxidation  2.667  pounds  of  oxygen.  Air,  by  volume,  is  composed 
of  about  21%  oxygen  and  79%  nitrogen,  and  by  weight,  of  23.15% 
oxygen  and  76.85%  nitrogen.  So  one  pound  of  oxygen  is  con- 
tained in  4.32  pounds  of  air.  It  therefore  takes  2.667X4.32= 
11.55  pounds  of  air  for  every  pound  of  carbon  burned.  At  ordi- 
nary room  temperatures  one  pound  of  air  occupies  about  13.4 
cubic  feet,  so  that  it  requires  theoretically  13.4X11.55=  155  cubic 
feet  of  air  for  the  complete  combustion  of  one  pound  of  carbon. 

The  hydrogen  in  the  hydrocarbons  of  the  coal  is  also  com- 
bustible. A  part  of  the  sulphur  present  may  be  combustible, 
but  it  is  usually  present  in  such  small  amounts  that  it  will  be 
omitted  in  our  present  computation.  As  explained  previously, 
not  all  of  the  hydrogen  content  of  the  coal  is  combustible,  since 
part  of  it  is  already  combined  with  oxygen  in  the  form  of  water. 
Therefore  the  available  hydrogen  may  be  expressed  as  (H  — O/8). 
Since  hydrogen  combines  with  oxygen  to  form  water,  in  the  ratio 
by  weight  of  1  to  8,  it  will  require  8  pounds  of  oxygen  to  burn 
each  pound  of  hydrogen.  Since  one  pound  of  oxygen  is  contained 
in  4.32  pounds  of  air,  it  will  take  8X4.32=34.6  pounds  of  air  to 
burn  a  pound  of  hydrogen.  Hence  the  total  weight  of  air  re- 
quired to  burn  a  pound  of  coal  to  CO2  and  H2O  is  theoretically 

11.55  C+34.6  (H-O/8), 

where  C,  H  and  O  have  the  same  meaning  as  in  §  13. 

Since  the  nitrogen  of  the  air  is  inert,  it  is  of  no  value  to  the 
combustion.  Since  it  passes  up  the  stack  at  a  higher  tempera- 
ture than  that  at  which  it  entered  the  furnace,  it  carries  away 
heat.  Any  less  air  than  the  theoretically  correct  amount  would 
result  in  the  formation  of  a  mixture  of  carbon  monoxide  and 


12  ENGINES   AND   BOILERS 

carbon  dioxide.  The  heat  liberated  by  the  formation  of  the  carbon 
monoxide  is  only  4450  B.t.u.  per  pound  of  carbon,  while  it  is 
nearly  14600  B.t.u.  for  the  formation  of  carbon  dioxide.  Hence  the 
production  of  carbon  monoxide  in  a  furnace  means  a  large  loss 
of  heat.  The  presence  of  carbon  monoxide  in  flue  gas  nearly 
always  indicates  a  large  amount  of  unburned  hydrocarbons  and 
hence  an  even  greater  loss  of  heat.  If  it  were  possible  to  so  dis- 
tribute the  air  that  it  all  came  in  close  contact  with  the  fuel,  and 
also  to  give  it  time  enough  to  combine  thoroughly  with  the  fuel, 
the  theoretical  amount  of  air  would  be  sufficient.  Under  actual 
furnace  conditions,  however,  it  is  found  that  50%  or  more  excess 
of  air  is  needed  to  give  complete  combustion  of  coal.  A  somewhat 
smaller  excess  is  needed  when  oil  is  used  as  a  fuel,  because  there 
is  better  distribution  of  the  air.  The  greater  the  amount  of  air 
passing  through  the  furnace,  the  greater  the  amount  of  heat  it 
will  carry  along  to  the  stack.  Hence  an  unnecessary  excess  of 
air  is  not  desirable,  and  leads  to  lessened  efficiency.  The  neces- 
sary excess  depends  upon  the  conditions  of  draft  and  fire  as  well 
as  upon  the  kind  of  fuel  and  the  type  of  furnace.  It  can  only 
be  determined  by  actual  test. 

15.  Composition  of  Flue  Gas.  —  As  just  explained,  an  excess 
of  air  is  needed  in  order  to  get  complete  combustion  of  the  fuel. 
If  it  were  possible  to  get  complete  combustion  without  this  excess, 
our  flue  gas  would  be  composed  chiefly  of  nitrogen,  carbon  dioxide, 
and  water  vapor.  Due  to  the  excess  of  air,  there  will  be  free 
oxygen  present  in  the  flue  gas.  If  there  is  an  insufficient  excess 
of  air  there  will  also  be  carbon  monoxide  and  probably  some 
hydrocarbons  present.  We  have  seen  that  the  presence  of  carbon 
monoxide  indicates  incomplete  combustion  and  therefore  low  fur- 
nace efficiency.  On  the  other  hand,  a  large  excess  of  air,  while 
it  may  give  complete  combustion,  gives  poor  furnace  efficiency 
because  the  air  will  carry  a  large  amount  of  heat  up  the  stack. 
It  is  a  matter  of  great  importance  that  just  the  right  excess  of 
air  be  admitted  to  the  furnace.  Since  it  is  difficult  to  measure 
directly  the  amount  of  air  entering  the  furnace,  an  easier  method 
is  used.  This  method  consists  in  analyzing  the  flue  gas  to  deter- 
mine the  amount  of  each  of  its  constituents.  From  this  analysis 
we  can  easily  compute  the  amount  of  air  entering  the  furnace. 
Knowing  the  composition  of  flue  gas,  we  can  regulate  the  amount 


FUEL  13 

of  air  entering  the  furnace  so  as  to  give  the  proper  excess  to  insure 
the  best  economy  of  operation. 

16.  Flue  Gas  Analysis.  —  There  are  various  types  of  apparatus 
on  the  market  for  making  the  analysis  of  flue  gas,  most  of  which 
are  modifications  of  the  apparatus  designed  by  ORSAT.     A  com- 
plete description  of  the  Orsat  apparatus  will  not  be  given  here, 
but  the  principle  of  its  operation  is  as  follows. 

A  sample  of  gas  is  taken  from  the  rear  of  the  furnace  or  between 
the  furnace  and  the  stack.  After  being  cooled  to  the  room  tem- 
perature, it  is  carefully  measured  by  volume  at  atmospheric 
pressure.  All  measurements  are  taken  at  room  temperature  and 
at  atmospheric  pressure.  This  known  volume  of  our  sample  is 
first  passed  a  few  times  through  a  solution  of  caustic  potash,  which 
absorbs  the  carbon  dioxide.  The  volume  is  measured  again,  and 
the  difference  between  the  new  volume  and  the  original  volume 
is  the  volume  of  the  carbon  dioxide  absorbed.  The  same  sample 
is  next  passed  several  times  through  a  solution  of  potassium  pyro- 
gallate,  which  absorbs  the  oxygen.  The  amount  of  oxygen  is 
determined  by  the  loss  in  volume,  as  before.  Next  the  sample 
is  passed  several  times  through  a  solution  of  acid  cuprous  chloride 
and  the  carbon  monoxide  removed,  and  its  amount  determined 
as  before.  The  amount  of  carbon  monoxide  is  usually  quite  small. 
The  remainder  of  the  sample  is  usually  assumed  to  be  nitrogen^ 

17.  Heat  Lost  in  Flue  Gas.  —  The  weight  of  flue  gas  per  pound 
of  fuel  burned  (assumed  carbon  and  ash)  may  be  computed  from 
the  formula, 


where 

W  =  weight  of  flue  gas  per  pound  of  fuel  burned. 
C  =  decimal  part  by  weight  of  total  carbon  in  fuel. 
N  =  percentage  by  volume  of  nitrogen  in  flue  gas. 
CC>2  =  percentage  by  volume  of  carbon  dioxide  in  flue  gas. 
CO  =  percentage  by  volume  of  carbon  monoxide  in  flue  gas. 

A  =  decimal  part  by  weight  of  ash  in  fuel  as  fired. 

Unless  the  ultimate  analysis  of  the  fuel  is  known,  the  weight 

of  carbon  in  the  volatile  matter  will  have  to  be  estimated  and 

added  to  the  weight  of  fixed  carbon  to  give  C  in  the  preceding 

formula.     Marks  has  published  a  chart  showing  the  approximate 


14 


ENGINES   AND   BOILERS 


relation  between  the  volatile  carbon  in  the  combustible  and  the 
total  volatile  matter  in  it.  With  the  aid  of  this  chart  (Fig.  4), 
the  value  for  C  in  the  preceding  formula  can  be  approximated 
from  the  proximate  analysis.  The  specific  heat  of  the  flue  gas 
is  usually  taken  as  .24;  and  the  heat  lost  per  pound  of  fuel 
burned  is  equal  to  the  product  of  the  specific  heat  of  flue  gas,  the 
weight  of  gas  per  pound  of  fuel,  and  the  difference  in  temperature 
between  the  leaving  flue  gas  and  the  entering  air. 

Chart  for  determining  the  Carbon  in  the  Volatile  Matter 

Marks 


ts 


20 


10  20  30  -9-0  SO  t 

Percent  of  Yo/otifa  Afatt-rr  jr?  ffye  Combust/hie 

FIG.  4 


EXAMPLE.  How  much  heat  is  carried  up  the  stack  by  the  dry  flue  gases, 
when  the  furnace  is  burning  coal  of  the  following  proximate  analysis?  Mois- 
ture  =  3%,  fixed  carbon  =  65%,  volatile  matter  =  26%,  and  ash=6%.  The 
analysis  of  flue  gases  gives:  CO2  =  10%,  O  =  8%,  and  CO  =  .5%.  The  stack 
temperature  is  500°  F.  and  the  temperature  of  the  air  entering  the  furnace 
is  80°  F. 

SOLUTION.    From  the  chart  of  Figure  4,  we  find  that  the  percent  ofjvoh 
tile  carbon  in  the  combustible  is  about  13.5,  which  corresponds 
fuel.     The  total  carbon  is  then  12.1+65  =  77.1%,  and  the  weighiToT  flue 
gases  per  pound  of  coal  is 


3.032  X.  771 


--°6)  =  19-1  Pounds. 


The  heat  carried  up  the  stack  by  the  dry  flue  gases  is  then  .24X19.1  (500  — 
80)  =  1925  B.t.u.  for  each  pound  of  coal  fired.  In  this  problem,  the  heat- 
value  of  a  pound  of  coal  found  from  the  proximate  analysis  is  13860  B.t.u. 
Hence  the  loss  is  1925/13860  =  13.9%  of  the  heat  available. 


FUEL  15 

Both  the  fuel  and  the  air  contain  moisture.  This  moisture  also  carries 
heat  up  the  stack,  since  it  is  a  superheated  vapor  upon  leaving  the  furnace. 

18.  Value  of  CO2  for  Best  Efficiency.  —  As  has  been  stated, 
the  efficiency  of  the  furnace  will  vary  with  the  excess  of  air  ad- 
mitted.   Since  the  percentage  of  CO2  also  varies  with  the  excess 
of  air,  we  see  that  an  indication  of  the  efficiency  .is  given  by  the 
CO2  reading.     Just  what  percentage  of  CO2  corresponds  to  the 
highest  operating  efficiency  depends  upon  such  factors  as  kind 
and   state   of  fuel,   stack  temperature,   etc.     After  the  proper 
amount  of  C02  for  best  efficiency  has  been  determined  under 
these  conditions,  the  CO2  reading  will  indicate  whether  or  not 
high  efficiency  is  being  obtained.    Since  the  determination  of  the 
CO2  is  a  comparatively  simple  operation,  it  is  an  excellent  way  to 
keep  check  on  operating  conditions.    Some  plants  go  so  far  as  to 
pay  their  firemen  on  the  basis  of  the  CO2  record.     In  general, 
high  CO2  means  high  efficiency,  unless  there  is  some  abnormal 
condition  such  as  too  much  CO.    The  CO  should  be  kept  as  near 
zero  as  possible. 

At  the  same  temperature  and  pressure,  CO2  occupies  the  same 
volume  as  the  oxygen  from  which  it  was  formed.  The  volume  of 
the  oxygen  in  the  air  is  21%.  Hence,  if  the  products  of  combus- 
tion are  cooled  down  to  the  temperature  of  the  entering  air,  the 
CO2  reading  would  be  21%  for  perfect  combustion  with  no  excess 
of  air,  assuming  the  fuel  to  be  carbon  and  ash.  In  practice,  the 
CO2  runs  17%  or  lower.  Even  15%  is  usually  considered  an  in- 
dication of  very  good  efficiency. 

19.  CO2  Recorders.  —  Automatic  devices  are  on  the  market 
that  will  analyze  and  record  the  amount  of  CO2  on  a  chart.    An 
analysis  is  made  every  few  minutes,  so  that  a  complete  record  is 
kept  of  the  operating  efficiency. 


CHAPTER  III 

STEAM 

20.  Introduction.  Definitions.  —  A  perfect  gas  may  be  at  any 
temperature  under  any  pressure.  For  instance,  air  may  be  placed 
under  a  certain  pressure  and  have  its  temperature  raised  or  low- 
ered by  the  addition  or  subtraction  of  heat.  On  the  other  hand, 
a  saturated  vapor,  such  as  steam,  can  exist  only  at  a  certain 
definite  temperature  for  each  particular  pressure.  Under  ordinary 
atmospheric  pressure,  saturated  steam  can  exist  only  at  a  temper- 
ature of  about  212°  F.  Under  an  absolute  pressure  of  100  pounds 
per  square  inch,  saturated  steam  will  be  at  a  temperature  of 
327.86°  F. 

Let  us  consider  a  case  in  which  a  pound  of  water  at  32°  F.  is 
placed  under  a  pressure  of,  say,  100  pounds  per  square  inch. 
The  containing  vessel  is  supposed  to  be  so  constructed  that  the 
pressure  remains  constant,  no  matter  what  change  of  volume 
takes  place.  Now  suppose  that  the  water  is  heated.  There  will 
be  little  change  in  volume,  but  there  will  be  a  rise  of  temperature 
of  approximately  one  degree  F.  for  each  B.t.u.  given  to  the  water. 
This  will  continue  until  we  have  added  298.5  B.t.u.  The  tem- 
perature will  then  be  327.86°  F.  A  further  addition  of  heat  up 
to  a  limit  will  not  cause  any  change  of  temperature,  but  will  effect 
a  change  in  the  physical  condition  of  the  water,  turning  it  to 
steam.  We  shall  need  to  apply  887.6  B.t.u.  to  effect  this  change 
completely.  We  have  now  added  a  total  of  298.5+887.6  =  1186.1 
B.t.u.,  and  have  converted  the  pound  of  water,  originally  at  32°  F., 
into  dry  saturated  steam  at  a  temperature  of  327.86°  F.,  and 
under  an  absolute  pressure  of  100  pounds  per  square  inch. 

Now  that  the  water  is  all  evaporated,  if  more  heat  be  added 
to  this  steam,  the  temperature  will  rise  at  the  rate  of  nearly  two 
degrees  per  B.t.u.  added.  We  now  have  superheated  steam. 

As  stated  previously,  the  volume  of  the  water  will  change  but 
little  until  the  boiling  point  is  reached.  The  space  occupied  by 
the  saturated  steam  will  be  4.432  cubic  feet.  This  is  many  times 
greater  than  the  space  formerly  occupied  by  the  water.  Part 
of  the  887.6  B.t.u.  that  was  used  to  evaporate  the  water  was 
evidently  used  to  cause  this  change  in  volume  under  the  pressure 

16 


STEAM 


17 


of  100  pounds  per  square  inch.  The  remainder  was  used  to  make 
the  physical  change  in  the  water,  to  increase  the  kinetic  energy 
of  its  atoms.  For  pressures  other  than  100  pounds  we  would 
have  values  different  from  those  given  above. 

Figure  5  represents  graphically  the  relation  between  the  tem- 
perature and  the  heat  added  to  a  pound  of  ice,  starting  at  zero 
degrees  F.  (with  the  assumption  that  the  pressure  is  constant). 
Upon  the  first  addition  of  heat,  the  temperature  of  the  ice  will 
rise  until  the  melting  point  is  reached.  Further  addition  of  heat 


1 

I 


O*  32 


Temperature 

FIG.  5 


causes  the  ice  to  melt.  This  occurs  without  a  change  of  tempera- 
ture. The  part  of  the  line  representing  the  melting  of  the  ice 
is  therefore  vertical.  When  the  ice  is  all  melted,  addition  of  more 
heat  causes  the  temperature  of  the  water  to  rise.  This  will  con- 
tinue until  the  boiling  point  is  reached.  That  part  of  the  line 
representing  evaporation  will  be  vertical  since  there  is  no  change 
of  temperature  during  that  period.  When  evaporation  is  com- 
plete, the  addition  cf  heat  again  causes  a  rise  in  temperature. 

The  amount  of  heat  necessary  to  raise  the  temperature  of  the 
water,  the  amount  of  heat  required  to  change  it  to  steam,  and 
also  the  volumes  of  steam  formed  under  different  pressures,  have 
been  determined  by  numerous  experiments  and  are  published 


18  ENGINES   AND   BOILERS 

under  the  name  of  steam  tables.  We  shall  use  in  our  work  the 
tables  prepared  by  C.  H.  Peabody.1  These  are  arranged  in  two 
ways.  In  Table  I,  the  various  absolute  pressures  at  which  water 
boils  are  given  for  each  degree  F.  from  32°  to  428°.  In  Table  II, 
the  various  temperatures  at  which  water  boils  are  given  for  each 
pound  per  square  inch  from  1  to  336.  The  values  are  arranged 
in  two  tables  not  because  they  are  different,  but  simply  as  a  con- 
venience in  their  use. 

In  Table  I,  the  first  column,  headed  t,  gives  the  temperature 
at  which  water  boils.  The  second  column,  headed  p,  gives  the 
absolute  pressure  under  which  it  must  be  in  order  that  it  boil 
at  the  temperature  given  in  the  first  column. 

The  third  column,  headed  q,  gives  the  heat  of  the  liquid,  which 
is  the  number  of  B.t.u.  necessary  to  change  the  temperature  of  one 
pound  of  water  from  32°  F.  to  the  temperature  given  in  the  first  column. 
This  does  not  mean  that  there  is  no  heat  in  the  water  at  32°. 
'The  heat  in  the  water  below  the  freezing  point  is  of  no  moment 
to  the  steam  engineer;  hence  it  is  chosen  as  the  arbitrary  starting 
point. 

Column  four,  headed  r,  gives  the  heat  of  vaporization,  which  is 
the  B.t.u.  necessary  to  evaporate  completely  a  pound  of  water  at  the 
temperature  and  pressure  given  in  the  first  and  second  columns. 
This  is  sometimes  called  the  latent  heat  of  evaporation.  The 
sum  of  the  heat  of  the  liquid  and  the  heat  of  vaporization  is 
called  the  total  heat. 

The  fifth  column,  headed  p,  is  that  part  of  the  heat  of  vaporiza- 
tion that  is  used  in  energizing  the  atoms  of  the  water  to  turn  it  to 
a  vapor;  it  is  called  the  heat  equivalent  of  internal  work. 

The  sixth  column,  headed  Apu,  is  the  rest  of  the  heat  of  vaporization, 
or  that  port  that  is  needed  to  do  the  work  of  increasing  the  volume, 
under  the  pressure  of  column  two;  it  is  called  the  heat  equivalent 
of  external  work. 

Columns  seven  and  eight  will  not  be  discussed  here.  Column 
nine,  headed  s,  gives  in  cubic  feet  the  specific  volume,  which  is 
the  volume  of  one  pound  of  dry  saturated  steam  under  the  pressure 
of  column  two. 

Column  ten  gives  the  reciprocals  of  the  values  found  in  column 
nine.  It  is  the  weight  of  one  cubic  foot  of  dry  saturated  steam  under 
the  pressure  of  column  two. 

i  C.  H.  PEABODY,  Steam  Tables. 


STEAM 


19 


Steam  generated  in  most  boilers  not  equipped  with  a  super- 
heater is  likely  to  carry  with  it,  when  leaving  through  the  outlet 
pipe,  a  small  amount  of  water  in  a  finely  divided  state  or  mist. 
Steam  containing  this  moisture  is  said  to  be  wet  steam.  The 

Chart  Showing  Specific  Heat  of  Superheated  Steam  Values  from 
Knoblauch  and  Jakob 


700 


zoo 


SO  /OO  /JO  ZOO  250 

Pressure  In  pounds  per  square  /nch  ffhso/ute 
FIG.  6. 


quality  of  wet  steam  is  expressed  in  percent.  If  in  a  hundred 
parts  by  weight  of  a  mixture  of  steam  and  water,  five  parts  by 
weight  are  moisture,  the  quality  of  the  mixture  is  said  to  be  95% 
and  the  priming  5%. 

As  long  as  steam  is  in  contact  with  water  it  will  remain  satu- 
rated, and  its  temperature  cannot  be  raised  under  constant  pres- 


20  ENGINES  AND   BOILERS 

sure.  If  it  is  conducted  away  from  the  water  and  led  to  a  super- 
heater, its  temperature  will  be  raised  by  the  addition  of  heat. 
It  is  then  superheated  steam.  The  amount  of  heat  necessary  to 
superheat  depends  upon  the  pressure  and  upon  the  degree  of 
superheat.  The  chart  of  Fig.  6  gives  the  specific  heat  of  super- 
heated steam  for  the  ranges  of  pressure  and  temperature  com- 
monly found  in  practice.  The  specific  heat  of  steam  varies  with 
both  temperature  and  pressure.  The  chart  gives  the  average 
values  of  specific  heat  as  the  steam  is  raised  from  the  temperature 
of  saturation  to  the  temperature  of  superheat. 

EXAMPLE  1.  How  much  heat  is  required  to  change  a  pound  of  water  at 
70°  F.  into  dry  saturated  steam  at  a  pressure  of  120  pounds  per  square  inch 
absolute? 

SOLUTION.  On  page  48  of  Peabody's  Steam  Tables,  we  find  that  the  heat 
of  the  liquid,  q,  for  120  pounds  pressure  is  312.3  B.t.u.  This  amount  of  heat 
would  bring  the  temperature  of  the  water  from  32°  F.  to  the  boiling  point. 
As  the  temperature  of  the  water  to  start  with  is  70°  (p.  36),  it  already  con- 
tains 38.1  B.t.u.  It  is  then  necessary  to  add  to  it  312.3-38.1=274.2  B.t.u. 
in  order  to  bring  it  to  the  boiling  point.  To  evaporate  the  water  requires 
the  heat  of  vaporization,  r,  at  120  pounds  (p.  48),  which  is  876.9.  Hence 
the  total  heat  required  to  bring  the  water  up  to  boiling  and  to  evaporate  it 
is  274.2+876.9  =  1151.1  B.t.u. 

EXAMPLE  2.  If,  in  Example  1,  the  quality  of  the  steam  formed  had  been 
97%,  how  much  heat  would  it  have  required? 

SOLUTION.  The  water  must  all  be  brought  to  the  boiling  point,  which 
takes  the  same  amount  of  heat  as  in  Example  1,  274.2  B.t.u.  As  the  quality 
is  97%,  only  .97X876.9  =  850.6  B.t.u.  are  needed  to  evaporate  the  water. 
Hence  the  total  amount  of  heat  required  is  274.2+850.6  =  1124.8  B.t.u. 

EXAMPLE  3.  Find  the  amount  of  heat  necessary  to  generate  the  pound 
of  steam  in  Example  1,  if  it  is  superheated  to  a  temperature  of  475°  F. 

SOLUTION.  To  generate  a  pound  of  dry  saturated  steam  under  the  con- 
ditions of  Example  1,  requires  1151.1  B.t.u.  The  temperature  correspond- 
ing to  120  pounds  pressure  is  341.3°  F.  The  superheat  is  then  475°  -341.3°  = 
133.7°.  From  the  chart  of  Fig.  6,  the  specific  heat  of  superheated  steam  is 
.537.  It  will  take  .537X133.7  =  71.8  B.t.u.  to  superheat  the  steam.  Hence 
the  total  heat  required  is  1151.1+71.8  =  1222.9  B.t.u. 

EXAMPLE  4.     Find  the  volume   of  the   pound  of  steam  in   Example   2. 

SOLUTION.  A  pound  of  dry  saturated  steam  at  120  pounds  pressure  occu- 
pies 3.723  cubic  feet.  The  quality  being  97%,  the  volume  occupied  by  the 
steam  is  .97X3.723  =  3.611  cubic  feet.  A  pound  of  water  occupies  .016 
cubic  feet,  and,  as  3%  of  the  pound  of  wet  steam  is  water,  the  volume  of  the 
water  is  .03  X  .016  =  .0005  cubic  feet.  Hence  the  total  volume  is  3.61 1 + .0005  = 
3.611  cubic  feet. 


STEAM 


21 


21.  The  Steam  Calorimeter.  —  In  making  tests  of  boilers  or 
engines  it  is  necessary  to  know  the  quality  of  steam  leaving  the 
one  or  entering  the  other.  A  steam  calorimeter  is  used  in  mak- 
ing this  determination.  Several  types  of  calorimeters  are  in  use. 
If  the  quality  of  steam  is  high  (between  94%  and  100%),  the 
throttling  type  is  usually  used.  When  properly  constructed  this 
calorimeter  is  sufficiently  accurate  for  ordinary  purposes. 


Figure  7  shows  a  throttling  calorimeter  attached  to  a  steam- 
pipe  A.  If  the  steam  is  saturated,  and  the  pressure  is  known 
from  a  pressure  gage  H,  its  temperature  may  be  determined  from 
the  steam  tables.  If  the  steam  in  A  is  superheated,  it  is  also 
necessary  to  take  its  temperature  by  means  of  a  thermometer, 
and  the  heat  contents  may  be  calculated  from  the  steam  tables. 
If  it  is  wet,  the  quality  must  be  known  to  find  its  heat  contents. 
The  tube  B  in  Fig.  7  is  a  sampling  tube  through  which  a  sample 
of  steam  is  taken  from  the  pipe  A.  This  sample  is  throttled  down 


22  ENGINES   AND    BOILERS 

in  pressure  at  C  from  the  pressure  in  A,  say  pi,  to  the  pressure 
in  the  calorimeter  G,  say  p%.  If  the  calorimeter  is  well  covered, 
but  little  heat  is  lost  by  radiation  and  the  heat  contents  in  one 
pound  of  steam  in  A  is  the  same  as  in  the  chamber  G.  The  pres- 
sure and  temperature  in  G  are  measured  by  the  gage  F  and  the 
thermometer  E,  and  the  heat  contents  are  computed  by  means 
of  the  steam  tables. 

Since  the  heat  content  is  the  same  per  pound  in  A  as  in  G, 
the  quality  in  the  former  may  be  computed  as  follows.  The  total 
heat  in  A  equals  q\-\-xri  where  x  is  the  quality  and  qi  and  r\  are 
the  heat  of  liquid  and  the  heat  of  vaporization  in  A,  respectively. 
If  the  calorimeter  is  working  properly,  the  steam  will  be  super- 
heated in  G,  and  its  heat  contents  will  be  equal  to 


where  q%  and  r%  are  the  heat  of  the  liquid  and  the  heat  of  vaporiza- 
tion in  G,  respectively,  £3  is  the  temperature  of  steam  in  G  as 
measured  by  the  thermometer  E,  and  k  is  the  temperature  of 
saturated  steam  at  the  pressure  p%.  The  term  .48  (£3  —  £2)  is  the 
heat  of  superheat  in  G,  since  .48  is  the  specific  heat  of  super- 
heated steam  at  low  pressure  and  temperature.  Then  we  have 


from  which  x  may  be  found. 

EXAMPLE.     Find  the  quality  of  steam  leaving  a  boiler  when  the  pressure 

is  165  pounds  gage.     The  gage  pressure  in  the  throttling  calorimeter  is  3 

founds,  the  temperature  is  265°  F.,  and  the  barometer  reading  is  29.6  inches. 

SOLUTION.     From  the  steam  tables,  q\  =345,  n  =851,  <?2  =  189,  r2  =  964,  and 

*2  =  221.     Then 

345+z851  =  189+964+.48(265-221), 

from  which  x  =  .975  or  97.5%. 


CHAPTER  IV 
BOILERS 

22.  Introduction.  —  The  stored  energy  in  fuels  is  utilized  by 
means  of  the  steam  engine  as  follows.  The  fuel  is  burned  in  a 
furnace,  resulting  in  a  mixture  of  heated  gases.  These  hot  gases 
pass  over  and  along  the  surface  of  a  boiler.  A  transfer  of  heat 
takes  place  through  those  boiler  surfaces  that  are  exposed  to 
contact  with  the  hot  gases  or  to  radiation  from  the  incandescent 
fuel  bed  on  the  one  side  and  water  or  steam  on  the  other  side. 
This  boiler  surface  is  called  heating  surface.  This  heat  is  con- 
ducted through  the  shell  of  the  boiler  and  is  spent  in  heating  and 
evaporating  the  water  contained  in  the  boiler.  The  steam  thus 
formed  is  conducted  from  the  boiler  to  the  engine  or  turbine, 
where  it  does  work  due  to  its  pressure  and  to  its  tendency  to 
expand. 

Practically  all  boilers  have  a  considerable  storage  of  heated 
water  and  steam.  This  water  and  steam  is  under  a  high  pres- 
sure and  would  increase  in  volume  hundreds  of  times  if  the  pres- 
sure were  removed.  A  sudden  release  of  this  pressure  causes 
an  explosion.  Many  lives  have  been  lost  and  a  great  amount  of 
property  destroyed  by  boiler  explosions.  Hence  the  consideration 
of  first  importance  in  a  boiler  is  its  safety.  Other  considerations 
are  its  first  cost,  its  life,  and  the  ease  with  which  it  may  be  cleaned 
and  repaired.  In  portable  boilers  and  marine  boilers,  weight  and 
the  space  occupied  are  of  great  importance. 

Since  the  purpose  of  the  heating  surface  is  to  conduct  heat 
from  the  furnace  to  the  water,  it  follows  that  the  conduction 
should  be  rapid  and  effective.  To  be  efficient,  the  boiler  must 
extract  a  large  proportion  of  the  heat  generated  in  the  furnace. 
The  surface  must  be  of  such  size  and  so  arranged  that  time  is 
allowed  to  render  this  transfer  as  complete  as  possible. 

Due  to  the  erosion  of  some  of  the  parts,  or  due  to  overheating 
consequent  on  the  formation  of  scale,  a  boiler  originally  fit  for  a 
certain  class  of  service  may  become  so  weakened  that  it  is  unsafe 
for  high  pressures.  Many  states  provide  for  an  inspection  of 
boilers  and  equipment  in  order  to  prevent  explosions.  The  vari- 
ous boiler  insurance  companies  also  make  periodic  inspection  of 
insured  boilers.  As  a  result  of  this  inspection,  the  inspector  sets 

23 


24  ENGINES   AND   BOILERS 

a  limit  to  the  pressure  that  the  boiler  may  carry.  The  fireman 
must  be  constantly  on  the  alert  to  detect  any  signs  of  a  develop- 
ing weakness. 

Much  of  the  water  used  in  boilers  will  deposit  scale  on  the  heat- 
ing surface  of  the  boiler.  This  scale  greatly  hinders  the  conduc- 
tion of  heat  to  the  water  and  may  even  become  thick  enough 
to  allow  the  metal  on  which  it  settles  to  become  overheated  to 
such  an  extent  that  it  gives  way  and  causes  an  explosion.  If  the 
scale  becomes  too  thick,  it  must  be  removed.  The  removal  and 
prevention  of  scale  will  be  considered  later. 

The  fire  side  of  the  heating  surface  in  many  boilers  will  collect 
soot  and  fine  ash.  To  maintain  efficient  steaming  these  surfaces 
must  be  kept  clean.  The  soot  should  be  swept  or  blown  from  the 
surface  as  fast  as  it  collects. 

In  the  construction  of  a  steam-boiler,  the  following  require- 
ments are  to  be  considered. 

(1)  Proper  materials  of  uniform  strength  and  reliability  must 
be  employed,  and  the  size  of  all  parts  must  be  so  designed  that 
a  sufficient  factor  of  safety  exists.     The  workmanship  must  be 
good. 

(2)  There  must  be  steam  space  and  water  capacity  such  that  a 
sudden  change  of  load  will  not  cause  an  undue  drop  in  steam 
pressure. 

(3)  There  must  be  a  sufficient  water  surface  to  allow  for  com- 
plete separation  of  the  steam  from  the  water.     Too  small  a  sur- 
face will  result  in  foaming. 

(4)  A  thorough  circulation  of  the  water  must  be  provided,  so 
that  a  uniform  temperature  is  maintained  throughout  the  boiler. 
Water  is  a  very  poor  conductor  of  heat,  and  it  is  therefore  essen- 
tial that  there  be  a  continuous  flow  over  the  heating  surface. 

(5)  Stresses  due  to  temperature  change  must  be  eliminated. 

(6)  In  so  far  as  possible,  all  joints  or  seams  must  be  protected 
from  the  direct  action  of  the  flame. 

(7)  Access  must  be  possible  to  all  parts  for  cleaning  and  repair. 

(8)  A  means  for  the  discharge  of  mud  or  sludge  that  is  left  by 
the  evaporation  of  the  water  must  be  provided. 

The  modern  steam  boiler  is  the  result  of  an  evolution  starting 
with  a  vessel  much  resembling  a  closed  kettle.  The  pressure  in  the 
early  boilers  was  nearly  atmospheric ;  hence  the  shape  was  not 
influenced  by  the  consideration  of  the  strength  of  the  boiler. 


BOILERS  25 

With  the  use  of  steam  under  pressure,  boilers  assumed  a  cylin- 
drical or  spherical  shape,  since  these  shapes  are  not  distorted  so 
much  by  internal  pressure.  The  simplest  boiler  is  cylindrical, 
with  hemispherical  ends. 

For  a  given  steam  pressure,  the  thickness  of  metal  required 
varies  directly  as  the  diameter.  Hence  heavy  plate  must  be  used 
for  boilers  of  considerable  size.  Moreover,  the  ratio  between  the 
heating  surface  and  the  volume  decreases  as  the  diameter  in- 
creases. Thus  it  is  seen  that  the  single  cylinder  is  suitable  for 
small  boilers  only. 

From  this  early  and  simple  type  of  boiler  the  development 
has  proceeded  along  two  distinct  lines.  First,  in  place  of  a  single 
large  cylinder,  several  smaller  ones  were  sometimes  used,  thereby 
decreasing  the  weight  of  metal  and  increasing  the  rate  of  steam- 
ing. Carrying  this  idea  still  further  results  in  a  large  number  of 
very  small  cylinders  or  tubes  filled  with  water  and  surrounded  by 
fire.  This  is  the  modern  water-tube  boiler. 

The  other  way  of  increasing  the  heating  surface  of  a  cylindrical 
boiler  is  to  run  smoke  flues  through  it.  The  earliest  types  con- 
tain one  or  two  large  flues.  The  modern  type  contains  a  large 
number  of  small  tubes  through  which  the  fire  or  the  products  of 
combustion  pass.  This  type  is  known  as  the  fire-tube  boiler. 

23.  Rated  Horsepower.  —  The  rating  of  a  boiler  is  usually 
based  on  its  heating  surface.     There  is  no  standard  for  this  rat- 
ing.    It  is  becoming  general  practice,  however,  to  rate  a  water- 
tube  boiler  on  a  basis  of  10  square  feet  of  heating  surface  per  boiler 
horsepower,  and  to  rate  a  fire-tube  boiler  at  11  or  12  square  feet 
per  boiler  horsepower.     Under  average  conditions  of  firing,  care, 
and  draft,  a  boiler  should  develop  good  economy  at  its  rated 
horsepower.    However,  many  boilers  are  able  to  carry  great  over- 
load and  still  give  excellent  efficiency.     Poor  efficiency  is  due  in 
general  more  to  overloading  the  furnace  than  to  increasing  the 
evaporation  from  the  boiler. 

24.  Heating  Surface.  —  It  has  been  customary  to  consider  as 
heating  surfaces  those  surfaces  which  have  water  on  one  side  and 
the  products  of  combustion  on  the  other  side.     No  distinction 
is  made  in  the  thickness  of  metal  in  different  parts  of  the  boiler, 
or  in  the  difference  in  temperature  of  the  gases  on  their  path  along 
the  heating  surface.     However,  great  accuracy  is  seldom  required 


26  ENGINES  AND   BOILERS 

in  computing  the  heating  surface  of  a  boiler.  In  calculating 
heating  surface,  the  inside  of  fire  tubes  and  the  outside  of  water 
tubes  is  used. 

25.  Rules  for  Finding  the  Heating  Surface.  — 

1.  Horizontal   Return-tubular   boilers.     Kent l  gives  the  fol- 
lowing rule: 

Take  the  dimensions  in  inches.  Multiply  two-thirds  of  the 
circumference  of  the  shell  by  its  length;  multiply  the  sum  of 
the  circumferences  of  all  the  tubes  by  their  common  length; 
to  the  sum  of  these  products  add  two-thirds  the  area  of  both 
tube  sheets;  from  this  sum  subtract  twice  the  area  of  all  the 
tubes;  divide  the  remainder  by  144  to  obtain  the  area  in  square 
feet. 

2.  Vertical  Tubular  boilers.     Kent2  gives  the  following  rule: 
Multiply  the  circumference  of  the  fire-box  (in  inches)  by  its 

height  above  the  grate;  multiply  the  combined  circumference 
of  all  the  tubes  by  their  length,  and  to  these  two  products  add 
the  area  of  the  lower  tube  sheet;  from  this  sum  subtract  the 
area  of  all  the  tubes,  and  divide  by  144 :  the  quotient  is  the 
number  of  square  feet  of  heating  surface. 

3.  General  rule.     The  U.  S.  Bureau  of  Mines  3  gives  the  fol- 
lowing rule: 

A  short  approximate  method  for  any  boiler  is  to  figure  the 
square  feet  of  heating  surface  in  the  tubes  and  divide  it  by  0.85 
for  a  return  tubular  boiler  or  by  0.95  for  a  water  tube  boiler. 
In  case  the  return  tubular  boiler  has  an  arch  over  the  top  for 
gas  passage,  giving  the  so-called  third  return,  it  is  necessary 
to  add  from  100  to  200  square  feet  to  the  result  to  obtain  the 
total  heating  surface. 

In  this  last  rule  the  heating  surface  in  fire-tube  boilers  is  figured 
from  the  outside  diameter  of  tubes. 

26.  Superheating  Surface.  —  In  modern  practice,  steam  is  often 
led  off  from  the  main  steam  space  and  taken  through  other  heat- 
ing coils.     Since  there  is  no  water  in  contact  with  this  steam,  it 

1  KENT,  Mechanical  Engineers'  Handbook,  1916  Edition,  p.  888. 

a  Ibid. 

8  BUREAU  OF  MINES,  U.  S.  DEPARTMENT  OF  THE  INTERIOR,  Bulletin  No.  40,  p.  9. 


BOILERS 


27 


will  be  superheated.  That  surface  which  has  this  superheated 
steam  on  one  side  and  fire  or  hot  gases  on  the  other  is  called  the 
superheating  surface. 

27.   Size  of  Boiler  Tubes.  —  The  outside  diameter  is  the  nom- 
inal diameter  in  boiler  tubes.     The  following  table  given  by  the 

5  team -pipe 
Cffnnecfion 


FIG.  8 

National  Tube  Works  gives  the  nominal  or  outside  diameter,  the 
inside  diameter,  and  the  thickness,  of  standard  lap-welded  boiler 
tubes. 

SIZE  IN   INCHES  OF  STANDARD  LAP-WELDED   TUBES 


External  Diameter 

1.0 
.810 
.095 
3.25 
3.010 
.120 

1.25 

1.060 
.095 
3.5 
3.260 
.120 

1.5 
1.310 
.095 
3.75 
3.510 
.120 

1.75 
1.560 
.095 
4.0 
3.732 
.134 

2.0 
1.810 
.095 
4.5 
4.232 
.134 

2.25 
2.060 
.095 
5.0 
4.704 
.148 

2.5 

2.282 
.109 
6.0 
5.670 
.165 

2.75 
2.532 
.109 
7.0 
6.670 
.165 

3.0 
2.782 
.109 
8.0 
7.675 
.165 

Internal  Diameter  
Standard  Thickness  
External  Diameter  
Internal  Diameter  
Standard  Thickness  

28  ENGINES   AND    BOILERS 

28.  Water-tube  Boiler.  Babcock  and  Wilcox  Type.  —  One  of 
the  most  common  forms  of  water-tube  boilers  is  the  Babcock  and 
Wilcox  type,  shown  in  Fig.  8.  Here  the  tubes  are  fastened  into 
headers,  which  in  turn  are  connected  by  other  tubes  to  a  forged 
steel  cross-box  that  is  riveted  to  the  steam  drum  above,  as  shown 
in  Fig.  9.  Headers  are  made  of  cast-iron  for  low-pressure  work, 
and  of  wrought  steel  for  high-pressure  work.  They  are  of  two 
types,  vertical  and  inclined.  In  the  latter  the  tubes  enter  the 
header  at  right  angles.  The  tubes  between  the  front  and  rear 
headers  are  inclined  so  that  a  circulation  of  water  is  insured. 
Opposite  the  end  of  each  tube  in  the  header  is  placed  a  hand- 
hole  to  permit  cleaning  and  access  to  the  tube.  These  holes  are 
closed  by  means  of  a  cap.  This  arrangement  also  allows  ease  in 
cleaning  the  scale  from  the  tubes,  as  the  caps  are  easily  removed 
and  a  cleaner  run  through  the  tube.  Clean-out  doors  are  placed 
in  the  setting  opposite  the  headers  to  give  easy  access  for  cleaning. 
The  grates  are  so  located,  and  fire-brick  partitions  or  baffles 
are  so  placed,  that  the  hot  gases  usually  pass  across  the  tubes 
^  three  times  on  their  way  to  the  stack.  The  mud 
|  |  drum  is  connected  to  the  lower  end  of  the  back 

header.  This  collects  the  sediment  that  is  formed 
during  the  evaporation  of  the  water.  This  sediment 
is  blown  off  from  time  to  time  through  the  blow-off 
pipe  shown  at  the  lower  right  hand  of  Fig.  8. 

The  feedwater  is  brought  in  at  the  front  of  the 
steam  drum,  and  is  so  delivered  that  its  velocity 
will  aid  in  circulation  of  the  water  in  the  boiler. 
The  circulation  is  back  through  the  steam  drum, 
down  the  tubes  to  the  back  header,  where  it  is  distributed  to  the 
inclined  tubes,  up  these  to  the  front  header,  and  thence  to  the 
front  of  the  steam  drum.  The  steam  is  taken  off  through  a  dry- 
pipe  located  at  the  top  of  the  steam  drum.  If  a  superheater  is 
attached,  as  shown  in  Fig.  8,  the  steam  is  led  from  the  dry-pipe  at 
the  top  of  the  steam  drum  into  and  through  the  superheater,  and 
then  up  to  the  steam  pipe  shown  at  the  top  of  the  steam  drum. 
Provision  is  made  for  flooding  the  superheater  during  the  period 
of  getting  up  steam.  This  keeps  the  surface  cool  enough  to 
prevent  its  burning  out.  The  boiler  is  carried  by  slings  from 
horizontal  girders  placed  above  it.  The  whole  boiler  is  surrounded 
by  a  smoke-tight  brick  setting. 


BOILERS 


29 


A  similar  form,  but  using  boiler  plate  headers  instead  of  cast 
or  forged  sectional  headers,  is  called  the  Heine  type. 

29.  Water-tube  Boiler.     Stirling  Type.  —  A  common  form  of 
water-tube  boiler  is-  shown  in  Fig.  10.     Here  banks  of  tubes  lead 


feed  water  inlet. 


Sfatmotrffef 


FIG.  10 


from  the  lower  drum,  or  mud  drum,  to  the  three  drums  above. 
The  water  is  taken  in  at  the  rear  upper  drum  and  the  steam  is 
drawn  off  at  the  top  of  the  middle  drum.  Arched  tubes  connect 
the  steam  space  of  the  upper  drums.  Access  to  the  inside  of  the 
drums,  for  the  purpose  of  expanding  the  tubes,  for  cleaning,  and 


30 


ENGINES   AND    BOILERS 


for  inspection,  is  had  by  means  of  a  manhole  located  at  the  end 
of  each  drum.  The  sediment  collects  and  is  blown  off  at  the 
bottom  of  the  mud  drum  through  the  blow  off  pipe  and  valve 
shown  at  the  lower  left  hand  corner  of  Fig.  10.  The  baffling  is 


J/easnov/kf 


connect/em 


\3mo/ie  out/rt 


FIG.  11 


so  arranged  that  the  products  of  combustion  are  forced  to  travel 
practically  the  entire  length  of  all  the  banks  of  tubes,  up  the  first 
bank,  down  the  second  and  up  the  third.  Clean-out  doors  for 
the  removal  of  soot  are  located  at  various  points  shown  in  Fig.  10. 

30.  Water-tube  Boiler.  Vertical  Type.  —  A  Wickes  vertical 
water-tube  boiler  is  shown  in  Fig.  11.  In  this  type  of  boiler  the 
tubes  are  straight  and  are  placed  vertically.  They  are  connected 
to  a  mud  drum  below  and  to  a  steam  drum  above.  The  products 


BOILERS  31 

of  combustion  pass  up  along  the  front  half  of  the  tubes  and  down 
the  back  half,  leaving  at  the  rear. 

31.  Fire-tube  Boiler.  Externally  Fired,  Return-tubular  Type. 
-  In  this  country,  where  moderate  pressures  are  wanted,  the 
common  type  of  fire-tube  boiler  is  the  externally  fired  return- 
tubular  boiler.  Figure  12  shows  the  construction  of  a  boiler  of 
this  type.  It  consists  of  a  cylindrical  shell,  made  of  steel  or 
wrought-iron  plates  rolled  into  a  cylindrical  shape  and  riveted 
together.  The  ends  or  heads  are  formed  from  flat  circular  plates 
flanged  around  the  outer  edge  and  riveted  to  the  cylindrical 
shell.  A  large  number  of  fire-tubes  extend  from  one  end-sheet 
to  the  other.  They  occupy  the  lower  two-thirds  of  the  shell, 
the  top  third  being  steam  space.  The  flat  plates  or  tube-sheets 
tend  to  bulge  outward  with  the  internal  pressure.  To  prevent 
this,  the  part  of  the  sheets  above  and  below  the  level  of  the  tubes 
must  be  stayed.  These  stays  are  of  two  kinds,  "through  stays," 
and  "diagonal  stays."  The  former  are  steel  rods  that  run  the 
entire  length  of  the  shell,  pierce  the  tube-sheets  and  hold  the 
sheets  in  by  means  of  nuts.  The  latter  run  from  each  tube-sheet 
diagonally  back  to  the  shell.  In  certain  types  of  large  boilers, 
some  of  the  tubes  are  made  heavier  and  are  threaded  on  the  ends. 
These  tubes  are  secured  to  the  end-sheets  by  means  of  nuts  on 
the  outside.  Such  tubes  are  called  stay-tubes.  In  general,  the 
tubes  act  as  stays,  and  that  part  of  the  ends  occupied  by  the  tubes 
needs  little  if  any  extra  staying.  The  tubes  are  expanded  into 
the  sheet  and  their  ends  are  beaded  over. 

The  feed-pipe  enters  the  front  end  of  the  boiler  just  below  the 
water  line,  and  the  steam  leaves  by  the  dry-pipe,  which  leads 
out  of  the  top  of  the  shell.  The  mud  is  blown  off  through  an 
outlet  at  the  bottom  near  the  rear.  If  the  feedwater  is  taken 
from  a  pond  or  stream  and  contains  much  vegetable  matter, 
there  should  be  a  surface  blow-off  to  skim  the  scum  that  forms 
on  top  of  the  water.  A  manhole  is  located  at  the  top  near  the 
rear  and  a  hand-hole  in  front  beneath  the  tubes.  The  grate  is 
put  beneath  the  front  end  of  the  shell.  The  products  of  com- 
bustion pass  over  the  bridge  wall,  back  along  the  bottom  of  the 
shell,  enter  the  tubes  from  the  rear,  pass  through  them  and  out 
of  the  uptake  at  the  front  of  the  boiler  or  through  a  flue  over 
the  top  of  the  boiler  to  an  uptake  at  the  rear. 


32 


ENGINES  AND   BOILERS 


1f 

r-J 

IT 

1   i! 

i  u  t,  

h 
n 
1  1 

, 

r 

K 

L 
r 

L. 

r--jp— 

ri'      ^      if 

.-,-  JJ                         LH» 

: 

1                        i 
i 
i 

-1- 

:d 

r 
P-. 

u 

r 

© 

•"Hi      ®      r 

.•ra'J                                    tfc 

~J 

o 

Jf 
i. 

—1 
H 

c    ; 

BOILERS 


The  shell  is  supported  by  brackets  which  are  riveted  to  the 
shell  and  rest  on  the  brickwork  of  the  setting.  The  rear  bearing 
is  fitted  with  rollers  to  allow  the  boiler  to  expand  without  crack- 
ing the  setting. 

32.  Fire-tube  Boiler.  Internally  Fired,  Return-tubular  Type. 
—  Instead  of  having  the  fire  outside  the  shell,  as  in  the  preced- 
ing boiler,  the  fire  is  sometimes  placed  inside  one  or  more  large 


Jftrahe  out/ft 


ooooooooc 
oooooooo 
ooooooooooc 
ooooooooo 
ooooooooooc 
oooooo 


FIG.  13a 

fire-flues  running  the  whole  length  of  boiler.  Figure  13  shows  a 
Springfield  boiler  of  this  type.  These  large  fire-flues  are  of  course 
subjected  to  external  pressure,  and  therefore  have  a  tendency  to 
collapse.  In  order  to  prevent  this,  they  must  be  stiffened  or 
stayed  in  some  way.  This  is  ordinarily  done  by  rolling  them  in 
corrugations.  If  the  flue  is  braced  in  such  a  manner  that  the 
metal  at  any  one  point  is  very  thick,  there  is  danger  of  its  being 
overheated,  since  the  gases  are  hottest  in  this  combustion  flue. 

The  fire  is  at  the  front  of  the  flue.  The  gases  pass  back  through 
the  flue  to  the  rear  of  the  boiler,  into  the  combustion  chamber. 
The  heating  surface  is  mostly  composed  of  the  comparatively 
thin  flue  and  tubes.  The  thick  outer  shell  is  not  subjected  to 


34 


ENGINES   AND    BOILERS 


the  high  temperatures  of  the  combustion  chamber,  as  in  the 
externally  fired  boiler. 

33.  Fire-tube  Boiler.  Scotch  Marine  Type.  —  While  water- 
tube  boilers  are  used  to  some  extent  in  marine  service,  the  stand- 
ard boiler  is  the  Scotch  marine  type.  This  is  very  similar  to  the 
internally  fired  boiler  described  in  §  32,  which  is  also  sometimes 
called  a  Scotch  marine  boiler.  It  commonly  has  three  combus- 


Snsafji  o/oen  t* 


& 


^    On/pifJ^TT^^^'    ' 


ta^X^VXK^VXVXV^V^IVi^S^> 


FIG.  13b 


tion  flues,  each  of  which  has  its  own  set  of  tubes.  In  marine 
practice  the  combustion  chamber  at  the  rear  of  the  flue  and  the 
tubes  is  internal,  there  being  a  water-leg  between  the  combustion 
chamber  and  the  rear  head.  There  is  a  combustion  chamber  for 
each  flue  and  its  set  of  tubes.  Since  the  surfaces  of  the  combus- 
tion chamber  are  flat,  they  must  be  stayed. 

These  boilers  are  large  in  diameter;  the  outer  shell  must  there- 
fore be  very  thick.  The  longitudinal  seams  are  usually  triple 
riveted,  with  two-strap  butt  joints.  The  Scotch  marine  boiler  is 
used  to  some  extent  on  land,  but  as  the  space  occupied  is  usually 
an  element  of  less  importance  here,  the  type  is  not  common. 


BOILERS 


35 


34.  Fire-tube  Boiler.     Vertical  Type.  —  A   boiler  often  used 
for  small  or  portable  plants  is  the  vertical  fire-tube  type  (Fig.  14). 
These  boilers  are  internally  fired,  the  fire  being  enclosed  in  the 
lower  part  of  the  shell,  and  surrounded  by  an  annular  water-ring 
or  water-leg.    The  lower  tube- 
sheet  is  placed  but   a  small 

distance  above  the  grate. 
Therefore  the  space  for  com- 
bustion is  very  limited.  The 
tubes  are  vertical  and  are  ex- 
panded into  the  lower  and  the 
upper  tube-sheets. 

,/  connecfioas 

35.  Fire-tube  Boiler.  Loco- 
motive Type.  —  The  type  of 
boiler    used   on    locomotives, 
and  also  often  used  on  portable 
plants,  is   shown  in  Fig.   15. 
In  this  boiler,  the  fire-box  is 
at  the  rear  end  of  the  shell, 
and  its  top  and  sides  are  water- 
heating  surfaces. 

Since  the  sheets  that  form 
the  water-legs  at  the  sides  and 
rear  of  the  fire-box  are  flat,  it 
is  necessary  to  stay  them  to 

prevent  distortion.  A  screw-stay  is  used  for  this  purpose;  it  con- 
sists of  a  threaded  bolt  screwed  through  the  parallel  plates.  The 
threads  on  the  center  part  of  these  stay-bolts  are  removed  so 
that  cracks  will  not  start  in  the  bolt  at  the  root  of  the  thread. 
On  what  are  called  safety  stays  a  small  hole  is  drilled  in  from 
the  end  so  that  a  cracked  bolt  will  leak  steam  and  give  warning. 

The  flat  or  arched  sheet  at  the  top  of  the  fire-box  is  called  the 
crown-sheet.  The  crown-sheet  is  stayed  in  various  ways,  some- 
times by  radial  stay-bolts  which  run  between  it  and  the  outer 
shell,  or  by  sling  stays,  which  are  girders  slung  from  the  outer 
shell.  Fire-tubes  extend  from  the  tube-sheet  at  the  front  of  the 
fire-box  to  the  tube-sheet  at  the  front  end  of  the  boiler. 

The  tubes  in  locomotive  boilers  are  smaller  than  in  the  types 
previously  described,  and  are  placed  as  close  together  as  good 


FIG.  14 


36 


ENGINES  AND   BOILERS 


BOILERS  37 

circulation  of  the  water  will  permit.  By  making  the  tubes  small 
and  numerous,  a  large  heating  surface  is  obtained.  Where  a 
superheater  is  used,  as  shown  in  Fig.  15,  some  of  the  tubes  are 
made  larger  and  contain  the  superheating  surface.  This  super- 
heating surface  is  formed  by  tubes  that  extend  into  the  fire-tubes 
from  the  front  end  of  the  boiler  and  run  to  within  a  short  dis- 
tance of  the  fire-box. 

The  outer  shell  extends  beyond  the  front  tube-plate  to  form 
a  smoke-box.  In  this  smoke-box  vertical  nozzles  are  located, 
through  which  the  exhaust  steam  from  the  engines  escapes.  This 
induces  a  strong  draft  that  allows  a  very  rapid  rate  of  combus- 
tion. The  rate  of  combustion  often  exceeds  100  pounds  of  coal 
per  square  foot  of  grate  surface  per  hour. 

Since  the  steaming  of  this  type  of  boiler  is  very  rapid,  the 
steam  is  taken  from  a  steam-dome  located  on  the  top  of  the 
shell.  This  allows  the  steam  to  be  taken  at  a  distance  from  the 
water  surface,  thereby  insuring  fairly  dry  steam.  The  throttle 
valve  for  the  engine  is  located  in  the  dome. 

In  smaller  locomotive  boilers,  the  fire-box  is  set  between  the 
rear  drivers.  In  the  larger  sizes,  this  arrangement  will  not  allow 
a  large  enough  grate  area,  and  so  the  fire-box  is  extended  later- 
ally over  low  trailer  wheels.  Manholes  and  hand-holes  are  em- 
ployed to  give  access  for  cleaning,  as  in  other  types  of  fire-tube 
boilers. 

36.  Superheaters.  —  During  the  past  few  years  superheated 
steam  has  come  into  quite  general  use,  especially  if  it  is  to  be 
used  in  steam  turbines.  The  amount  of  superheat  used  is  gener- 
ally not  large,  usually  between  100°  and  200°  Fahrenheit.  The 
advantages  gained  by  the  use  of  superheated  over  saturated 
steam  will  be  considered  later. 

There  are  two  types  of  superheater.  One  type  is  independ- 
ently fired.  The  other  is  formed  by  the  addition  of  some  super- 
heating surface  to  the  main  boiler.  The  latter  is  the  more  com- 
mon form.  In  this  type  the  steam  is  taken  from  the  steam  space 
and  led  through  superheating  coils.  Often  provision  is  made  for 
the  flooding  of  these  coils  during  the  period  in  which  steam  is 
being  raised,  in  order  to  prevent  the  coils  from  burning  out.  The 
boiler  shown  in  Fig.  8  has  a  common  form  of  superheater  at- 
tached. 


38  ENGINES   AND    BOILERS 

37.  Horsepower  of  Boilers. — As  explained  previously,  boilers 
are  generally  rated  by  the  manufacturer  on  the  amount  of  their 
heating  surface.  The  rate  at  which  a  boiler  is  working,  how- 
ever, must  be  determined  from  a  consideration  of  the  amount  of 
steam  that  is  being  generated. 

The  amount  of  heat  necessary  to  evaporate  a  given  quantity 
of  water  varies  with  the  temperature  of  the  feedwater,  with  the 
pressure  at  which  the  steam  is  formed,  and  with  the  quality  of 
the  steam  produced.  Hence  it  is  desirable  that  there  be  a  stand- 
ard of  temperature  and  pressure  at  which  we  can  find  the  equiv- 
alent amount  of  water  evaporated,  using  the  same  amount  of 
heat  as  is  used  under  the  actual  conditions  of  temperature  and 
pressure.  The  conditions  of  temperature  and  pressure  set  by 
the  A.  S.  M.  E.1  as  a  standard  are  212°  F.  and  14.7  pounds  per 
square  inch  absolute.  The  equivalent  evaporation  is  then  the 
amount  of  water  that  would  be  evaporated  from  and  at  212°  F. 
if  the  same  amount  of  heat  were  used  in  its  evaporation  as  is 
used  in  the  evaporation  under  the  actual  working  conditions. 
From  the  steam  tables,  the  B.t.u.  required  to  evaporate  one 
pound  of  water  from  and  at  212°  F.  is  seen  to  be  969.9  B.t.u. 
This  is  the  unit  of  evaporation. 

At  the  time  when  it  first  became  necessary  to  rate  boilers,  a 
good  engine  used  about  30  pounds  of  steam  per  horsepower  per 
hour  at  a  pressure  of  70  pounds  gage.  The  judges  at  the  Cen- 
tennial Exposition  in  1876  awarded  prizes  using  the  following 
unit  as  a  standard.  A  one-horsepower  boiler  is  one  that  will  evapo- 
rate 30  pounds  of  water  per  hour  from  feedwater  at  100°  F.  into 
steam  at  70  pounds  pressure  by  gage. 

The  A.  S.  M.  E.  has  since  adopted  an  equivalent  standard  and 
defines  a  boiler  horsepower  to  be  the  evaporation  of  34.5  pounds 
of  water  per  hour  at  212°  F.  into  steam  at  212°  F.  and  at  a  pressure 
of  14.7  pounds  absolute  pressure.  As  the  heat  of  vaporization 
at  212°  F.  or  14.7  pounds  absolute  is  969.7  B.t.u.,  it  therefore 
takes  969.7X34.5  B.t.u.  per  hour  for  one  boiler  horsepower. 

To  determine  the  horsepower  at  which  a  boiler  is  working,  we 
must  therefore  first  find  how  many  B.t.u.  are  used  to  generate 
one  pound  of  steam  under  the  given  conditions  of  steam  pressure 
and  temperature  of  feedwater.  The  number  of  pounds  of  water 
evaporated  per  hour  multiplied  by  the  number  of  B.t.u.  to  gen- 

iA.  S.  M.  E.  POWER  TEST  CODE,  Edition  of  1915,  Table  4,  §21. 


BOILERS  39 

erate  one  pound  of  steam  under  the  conditions  will  give  the  num- 
ber of  B.t.u.  used  by  the  boiler  per  hour.  This  product  divided 
by  969.7X34.5  will  give  us  the  horsepower  of  the  boiler. 

EXAMPLE.  It  is  required  to  find  the  horsepower  of  a  boiler  working  under 
the  following  conditions: 

Steam  pressure  =  115  pounds  gage. 
Temperature  of  feedwater  =  65°  F. 
Quality  of  steam  =  98%  (i.e.,  2%  priming). 
Water  fed  to  boiler  per  hour  =  3640  pounds. 
SOLUTION.     From  the  steam  tables  it  is  seen  that 
The  heat  of  the  liquid  at  115  Ib.  gage  (129.7  Ib.  abs.)  =  318.4  B.t.u. 
The  heat  of  vaporization  at  115  pounds  gage  =872.3 

The  heat  of  the  liquid  at  65°  F.  =33.1  B.t.u.  The  heat  required  to  evapo- 
rate one  pound  of  water  under  the  above  conditions  is  318. 4  +  . 98X872. 3  — 
33.1  =  1140  B.t.u.,  and  the  B.t.u.  used  per  hour  is  1140X3640  =  4150000. 
Hence  the  horsepower  of  the  boiler  is  4150000/(34.5X969.7)  =  124  h.p. 

38.  Factor  of  Evaporation.  —  Since  it  takes  969.7  B.t.u.  to 
evaporate  a  pound  of  water  from  and  at  212°  F.,  and  since  it 
takes  more  (1140  B.t.u.  in  the  previous  example)  to  evaporate 
a  pound  of  water  under  the  actual  conditions  that  exist  in  the 
boiler,  a  certain  ratio  exists  between  these  amounts.    The  factor 
of  evaporation  is  the  ratio  of  the  amount  of  heat  required  to  evap- 
orate a  pound  of  water  under  actual  conditions  to  the  amount  re- 
quired to  evaporate  a  pound  from  and  at  212°  F.     In  the  previous 
example,  the  factor  of  evaporation  was   1140/969.7  =  1.176.     If 
the  factor  of  evaporation  is  known,  the  equivalent  evaporation  is 
found  by  multiplying  the  actual  evaporation  by  this  factor.1 

By  this  method,  the  heat  used  to  raise  the  temperature  of  the 
moisture  in  the  steam  from  the  temperature  of  the  feed  water  to 
that  of  the  steam  is  not  considered  in  computing  the  factor  of 
evaporation.  In  most  cases  the  difference  in  results  due  to  this 
omission  is  very  small. 

39.  Efficiency  of  Boilers.  —  Usually  speaking,  the  efficiency  of 
anything  is  the  ratio  of  what  is  got  out  to  what  is  put  in;  output 
and  input  being   measured  in  like  units.     For  boilers,  the  term 
efficiency   means  the  ratio  of  the  number  of  B.t.u.  in  the  steam 
generated  to  the  number  of  B.t.u.  available  in  the  coal  fired.     Boiler 
efficiency  is  usually  expressed  in  percent. 

1  In  the  A.  S.  M.  E.  POWER  TEST  CODE  the  "Mean  B.t.u."  and  steam  tables  by  MARKS 
and  DAVIS  are  used,  thereby  giving  970.4  B.t.u.  instead  of  969.7  B.t.u.  as  used  above  for  heat 
required  to  evaporate  a  pound  of  water  from  and  at  212°  F.  See  POWER  TEST  CODE,  Edition 
of  1915,  pp.  28  and  47. 


40  ENGINES   AND    BOILERS 

The  fact  that  the  combined  efficiency  of  a  boiler,  furnace,  and 
grate  is  not  100%  is  due  to  several  losses.  These  losses  are  due 
to  the  following  causes. 

(1)  A  part  of  the  fuel  may  drop  through  the  grate  and  be  lost 
in  the  ash. 

(2)  Heat  is  lost  up  the  stack.     There  are  several  sources  of  this 
loss,  and  to  them  is  due  the  greatest  loss  in  efficiency.      First, 
unburned  particles  of  solid  fuel  are  often  carried  from  the  fur- 
nace.    The  amount  depends  upon  the  draft  and  the  kind  of  fuel. 
In  locomotives,  with  their  high  draft  and  with  a  fine  fuel,  this 
loss  may  be  considerable.     Second,  there  is  loss  due  to  the  un- 
burned or  partially  burned  hydrocarbons.     Black  smoke  is  caused 
by  the  incomplete  burning  of  some  of  the  hydrocarbons.     Third, 
heat  is  carried  away  by  the  excess  air  and  the  inert  nitrogen 
which  have  been  heated,  and  by  the  hot  products  of  combustion. 
Fourth,  heat  is  required   to  evaporate   and   to   superheat  the 
moisture  in  the  fuel  and  in  the  air.     Fifth,  there  may  be  loss 
due  to  the  burning  of  the  carbon  to  carbon  monoxide  instead  of 
to  carbon  dioxide. 

(3)  Heat  is  lost  by  radiation  from  the  furnace  and  from  the 
boiler. 

It  is  very  difficult  to  separate  all  these  losses  and  the  attempt 
is  seldom  made.  It  must  be  remembered  that  what  is  often 
called  boiler  efficiency  is  really  the  combined  efficiency  of  grate, 
furnace,  and  boiler. 

Under  the  most  favorable  conditions,  using  coal  as  a  fuel, 
efficiencies  of  over  80%  have  been  attained.  Under  ordinary 
conditions  of  operation,  efficiencies  vary  from  80%  to  less  than 
50%.  The  efficiency  may  sometimes  exceed  80%  when  underfeed 
stokers,  described  later,  are  used.  When  oil  is  used  as  a  fuel, 
higher  efficiencies  may  be  attained,  due  in  part  to  the  better  mix- 
ing of  the  air  and  the  fuel. 

EXAMPLE.  It  is  required  to  find  the  combined  efficiency  of  a  boiler,  fur- 
nace, and  grate,  working  under  the  following  conditions: 

Steam  pressure  =  127  pounds  gage. 

Superheat  =  190°  F. 

Temperature  of  feedwater  =  180°  F. 

Water  fed  to  boiler  per  hour  =  8750  pounds. 

Coal  fired  per  hour  =  1 160  pounds. 

B.t.u.  per  pound  of  coal  as  fired  =  11540  B.t.u. 


BOILERS  41 

SOLUTION.  The  B.t.u.  required  to  generate  one  pound  of  steam  under 
the  above  conditions  is  seen  to  be 

325.4+866.8  - 148+.55  X 190  =  1148.7  B.t.u. 

The  total  B.t.u.  used  in  the  generation  of  steam  per  hour  then  is  8750  X  1148.7  = 
10051000  B.t.u.  The  total  B.t.u.  in  coal  fired  per  hour  is  1160X11540  = 
13386000  B.t.u.  Hence  the  efficiency  is  10051000/13386000  =  .752  or  75.2%. 

40.  A.  S.  M.  E.  Boiler  Test  Code.1  —  In  reporting  the  results 
of  a  steam-boiler  test  it  is  well  to  put  them  in  the  form  prescribed 
by  the  A.  S.  M.  E.  This  form  is  as  follows. 

DATA  AND  RESULTS  OF   EVAPORATIVE  TEST 
CODE  OF   1915 


(1)  Test  of boiler  located. 

To  determine 

Test  conducted  by 


DIMENSIONS 

(2)  Number  and  kind  of  boilers 

(3)  Kind  of  furnace 

(4)  Grate  surface  (width length ) sq.  ft. 

(a)  Approximate  width  of  air  openings  in  grate in. 

(6)  Percentage  of  area  of  air  openings  to  grate  surface per  cent 

(5)  Water  heating  surface sq.  ft. 

(6)  Superheating  surface sq.  f t. 

(7)  Total  heating  surface sq.  f  t. 

(a)  Ratio  of  water  heating  surface  to  grate  surface 

(6)  Ratio  of  total  heating  surface  to  grate  surface 

(c)  Ratio  of  minimum  draft  area  to  grate  surface 

(d)  Volume  of  combustion  space  between  grate  and  heating  surface 

cu.  ft. 

(e)  Distance  from  center  of  grate  to  nearest  heating  surface ft. 

DATE,  DURATION,  ETC. 

(8)  Date 

(9)  Duration hr. 

(10)  Kind  and  size  of  coal 

AVERAGE  PRESSURES,  TEMPERATURES,  ETC. 

(11)  Steam  pressure  by  gage Ib.  per  sq.  in. 

(a)  Barometric    pressure in.    of    mercury 

(12)  Temperature  of  steam,  if  superheated deg. 

(a)  Normal  temperature  of  saturated  steam deg. 

(13)  Temperature  of  feedwater  entering  boiler deg. 

(a)  Temperature  of  feedwater  entering  economizer deg. 

(6)  Increase  of  temperature  of  water  due  to  economizer deg. 

i  A.  S.  M.  E.  POWER  TEST  CODE,  Edition  of  1915,  p.  51. 


42  ENGINES   AND   BOILERS 

(14)  Temperature  of  escaping  gases  leaving  boiler deg. 

(a)  Temperature  of  gases  leaving  economizer deg. 

(6)  Decrease  of  temperature  of  gases  due  to  economizer deg. 

(c)   Temperature  of  furnace deg. 

(15)  Force  of  draft  between  damper  and  boiler in.  of  water 

(a)  Draft  in  main  flue  near  boiler in.  of  water 

(6)  Draft  in  flue  between  economizer  and  chimney in.  of  water 

(c)  Draft  in  furnace in.  of  water 

(d)  Draft  or  blast  in  ash-pit in.  of  water 

(16)  State  of  weather 

(a)  Temperature  of  external  air deg. 

(6)  Temperature  of  air  entering  ash-pit deg. 

(c)  Relative  humidity  of  air  entering  ash-pit per  cent 

QUALITY  OF  STEAM 

(17)  Percentage  of  moisture  in  steam  or  number  of  degrees  of  superheating 

per  cent  or  deg. 

(18)  Factor  of  correction  for  quality  of  steam 

TOTAL  QUANTITIES 

(19)  Total  weight  of  coal  as  fired Ib. 

C 20)  Percentage  of  moisture  in  coal  as  fired per  cent 

(21)  Total  weight  of  dry  coal  (Item  19X 

(22)  Ash,  clinkers,  and  refuse  (dry) 

(a)  Withdrawn  from  furnace  and  ash-pit Ib. 

(6)  Withdrawn  from  tubes,  flues  and  combustion  chamber Ib. 

(c)   Blown  away  with  gases Ib. 

(d}  Total Ib. 

(e)  Weight  of  clinkers  contained  in  total  ash Ib. 

(23)  Total  combustible  burned  (Item  21  —Item  22d) Ib. 

(24)  Percentage  of  ash  and  refuse  based  on  dry  coal per  cent 

(25)  Total  weight  of  water  fed  to  boiler Ib. 

(26)  Total  water  evaporated,  corrected  for  quality  of  steam  (Item  25  X  Item 

18) Ib. 

(27)  Factor  of  evaporation  based  on  temperature  of  water  entering  boiler.  . . . 

(28)  Total  equivalent  evaporation  from  and  at  212  degrees  (Item  26Xltem 

27) Ib. 

HOURLY  QUANTITIES  AND  RATES 

(29)  Dry  coal  per  hour Ib. 

(30)  Dry  coal  per  square  foot  of  grate  surface  per  hour Ib. 

(31)  Water  evaporated  per  hour,  corrected  for  quality  of  steam Ib. 

(32)  Equivalent  evaporation  per  hour  from  and  at  212° Ib. 

(33)  Equivalent  evaporation  per  hour  from  and  at  212°  and  per  square  foot 

of  water  heating  surface Ib. 

CAPACITY 

(34)  Evaporation  per  hour  from  and  at  212°  (Same  as  Item  32) Ib. 

(a)  Boiler  horsepower  developed  (Item  34-5-34^) bl.-h.p. 


BOILERS  43 

(35)  Rated  capacity  per  hour,  from  and  at  212°  ........................  lb. 

(a)  Rated  boiler  horsepower  ..............................  bl.-h.p. 

(36)  Percentage  of  rated  capacity  developed  ......................  per  cent 

ECONOMY 

(37)  Water  fed  per  pound  of  coal  as  fired  (Item  25-r-Item  19)  ...........  lb. 

(38)  Water  evaporated  per  pound  of  dry  coal  (Item  26  -^  Item  21)  ........  lb. 

(39)  Equivalent  evaporation  from  and  at  212°  per  pound  of  coal  as  fired 

(Item  28  4-  Item  19)  ..........................................  lb. 

(40)  Equivalent  evaporation  from  and  at  212°  per  pound  of  dry  coal  (Item 

28^-  Item  21)  ...............................................  lb. 

(41)  Equivalent  evaporation  from  and  at  212°  per  pound  of  combustible 

(Item  28  -f-  Item  23)  ..........................................  lb. 

EFFICIENCY 

(42)  Calorific  value  of  1  pound  of  dry  coal  by  calorimeter  ............  B.t.u. 

(a)  Calorific  value  of  1  pound  of  dry  coal  by  analysis  .........  B.t.u. 

(43)  Calorific  value  of  1  pound  of  combustible  by  calorimeter  .........  B.t.u. 

(a)  Calorific  value  of  1  pound  of  combustible  by  analysis  ......  B.t.u. 

(44)  Efficiency  of  boiler,  furnace  and  grate 

/  Item40X970.4\ 

(100X    -TtoZlar-)  ................  percent 

(45)  Efficiency  based  on  combustible 

/in       Item4lX970.4\ 

(100X-     Item  43       J  ...............  P6F  C6nt 

COST  OF  EVAPORATION 

(46)  Cost  of  coal  per  ton  of  ......  pounds  delivered  in  boiler  room  ......... 


................  dollars 

(47)  Cost  of  coal  required  for  evaporating  1000  pounds  of  water  under  ob- 

served conditions  .........................................  dollars 

(48)  Cost  of  coal  required  for  evaporating  1000  pounds  of  water  from  and 

at  212°  ....................................................  dollars 

SMOKE  DATA 

(49)  Percentage  of  smoke  as  observed  ............................  per  cent 

(a)  Weight  of  soot  per  hour  obtained  from  smoke  meter  .....  per  cent 

FIRING  DATA 

(50)  Kind  of  firing,  whether  spreading,  alternate  or  coking  ................ 

(a)  Average  thickness  of  fire  ..................................  in. 

(6)  Average  intervals  between  firings  for  each  furnace  during  time 
when  fires  are  in  normal  condition  ......................  min. 

(c)  Average  interval  between  times  of  leveling  or  breaking  up  ....... 

.  .  min. 


44 


ENGINES   AND   BOILERS 


(51)  Analysis  of  dry  gases  by  volume 

(a)  Carbon  dioxide  (CCh) per  cent 

(b)  Oxygen  (O) per  cent 

(c)  Carbon  monoxide  (CO) per  cent 

(d)  Hydrogen  and  hydrocarbons per  cent 

(e)  Nitrogen,  by  difference  (N) per  cent 

(52)  Proximate  analysis  of  coal 

As  fired  Dry  coal  Combustible 

(a)  Moisture 

(6)  Volatile  matter 

(c)  Fixed  carbon. ...          

(d)  Ash 

100  per  cent         100  per  cent         100  per  cent 

(e)  Sulphur,  separately  determined  referred  to  dry  coal per  cent 

(53)  Ultimate  analysis  of  dry  coal 

(a)  Carbon  (C) per  cent 

(b)  Hydrogen  (H) per  cent 

(c)  Oxygen  (O) per  cent 

(d)  Nitrogen  (N) per  cent 

(e)  Sulphur  (S) per  cent 

(/)   Ash per  cent 

100  per  cent 

(54)  Analysis  of  ash  and  refuse,  etc. 

(a)  Volatile  matter per  cent 

(6)  Carbon per  cent 

(c)  Earthy  matter per  cent 

100  per  cent 
(d}  Sulphur,  separately  determined per  cent 

(d)  Fusing  temperature  of  ash deg. 

(55)  Heat  balance  based  on  dry  coal . 


Dry  Coal 


B.t.u. 


Percent 


(a)  Heat  absorbed  by  the  boiler  (Item  40X970.4) .  .  . 
(6)  Loss  due  to  evaporation  of  moisture  in  coal 

(c)  Loss  due  to  heat  carried  away  by  steam   formed 

by  the  burning  of  hydrogen 

(d)  Loss  due  to  heat  carried  away  in  the  dry  flue  gases 

(e)  Loss  due  to  carbon  monoxide 

OLoss  due  to  combustible  in  ash  and  refuse 
i  Loss  due  to  heating  moisture  in  air 

(h)  Loss  due  to  unconsumed  hydrogen  and  hydrocar- 
bons, to  radiation  and  unaccounted  for .  . 


(i)   Total   calorific   value   of   1   pound   of  dry  coal 
(Item  42) 


100 


CHAPTER   V 
BOILER  ACCESSORIES  AND   AUXILIARIES 

41.  Grates.  —  Grates  are  used  to  support  the  fuel  in  a  furnace. 
Most  grates  are  made  of  cast  iron,  which  is  cheap  and  less  liable 
than  other  convenient  materials  to  be  distorted  or  twisted  under 
the  high  temperatures  to  which  it  is  subjected.     The  grate  must 
be  strong  enough  to  support  the  load  placed  upon  it,  and  it  must 
be  of  such  a  form  that  sections  can  easily  be  replaced  when  broken 
or  burned  out.     It  must  have  sufficient  opening  for  the  admis- 
sion of  air  to  the  fuel.     The  openings  or  air  spaces  depend  upon 
the  kind  of  fuel  used.     The  combined  area  of  the  openings  will 
usually  be  from  30  to  50  percent  of  the  total  area. 

The  area  of  the  grate  depends  upon  the  amount  of  coal  to  be 
burned  and  the  rate  of  combustion.  Under  natural  or  chimney 
draft,  from  10  to  25  pounds  of  coal  can  be  burned  per  square 
foot  of  grate  surface  per  hour.  Under  forced  draft,  from  40  to 
130  pounds  of  coal  may  be  burned  per  hour.  If  hand  firing  is 
employed,  the  grate  must  not  be  longer  than  the  distance  the 
fireman  can  throw  the  coal  accurately  (six  or  seven  feet) .  Depend- 
ing upon  the  fuel,  draft,  and  economy  of  the  boiler,  the  equivalent 
evaporation  per  pound  of  coal  will  vary  from  5  to  12. 

Various  forms  of  grates  are  used.  For  hand  firing,  plain  grates 
and  shaking  or  dumping  grates  are  used.  The  plain  grate  is 
harder  to  keep  clean  than  a  dumping  grate.  Moreover,  it  is 
necessary  to  keep  the  fire-doors  open  while  the  cleaning  is  in 
process.  The  grates  used  in  mechanical  stokers  are  of  various 
types;  some  are  stationary,  others  traveling  and  rocking.  Occa- 
sionally grates  are  water-cooled,  to  prevent  their  burning  out. 
Since  this  water  is  led  to  the  boiler  after  becoming  heated  in  the 
grate,  the  boiler  capacity  is  increased,  but  in  most  cases  the  extra 
care  and  cost  are  prohibitive. 

42.  The  Plain  Grate.  —  The  grate  bars  shown  in  Figs.  16  and 
16a  are  of  the  stationary  type.     These  grates  are  cast  in  small 
sections  so  that  a  section  may  be  easily  and  quickly  replaced 
when  it  is  burned  out.     The  size  of  the  openings  in  the  bars  is 
governed  by  the  size  and  kind  of  coal  that  is  to  be  burned.     If 

45 


46 


ENGINES    AND    BOILERS 


anthracite  coal  is  used,  the  openings  are  small.     If  the  coal  is 
bituminous,  and  if  it  cakes,  the  openings  should  be  made  large. 

43.  The  Rocking  Grate.  —  A  form  of  rocking  grate  is  shown  in 
Fig.  17.     The  bars  are  supported  on  pivots,  and  are  dumped  or 
rocked  by  means  of  a  lever  from  the  front  of  the  furnace.     Only  the 
largest  clinker  need  be  removed  from  the  top,  since  the  rocking 
action  of  the  bars  breaks  up  most  of  the  clinker  that  is  formed. 
In  case  a  strong  draft  is  used,  as  in  the  locomotive,  this  type  of 
grate  is  usually  used  in  order  to  keep  a  clean  fire,  such  as  is  required 
with  a  high  rate  of  combustion. 

44.  Mechanical    Stokers.  —  The   first    cost   of    a   mechanical 
stoker   is   greater   than   the   equipment   for   hand   firing,   but  it 


Herat/on  \g 


cfioi?  at  4-3  ' 


FIG.  16 


FIG.  16a 


tr        rr 


requires  less  labor  and  attendance  in  its  operation.  A  cheaper 
grade  of  fuel  can  be  used,  a  higher  efficiency  attained,  and  less 
smoke  is  formed  than  is  usual  with  hand  firing.  In  a  fair-sized 
or  large  plant,  it  is  usually  better  economy  to  use  some  form  of 
mechanical  stoker.  There  are  many  forms  of  stokers  in  use. 

45.  The  Chain  Grate  Stoker.  —  Where  a  low-grade  fuel  is 
used,  as  is  often  the  case  in  the  middle  west  and  in  the  central 
states,  the  chain  grate  (Fig.  18)  is  extensively  used.  The  grate  is 
composed  of  a  large  number  of  short  links,  forming  an  endless 
chain.  This  chain  runs  over  front  and  rear  sprockets.  Power  is 
used  to  drive  one  of  these  sprockets,  causing  the  whole  chain 
to  revolve  slowly  at  a  speech  which  is  regulated  by  a  suitable 
mechanism.  The  whole  grate  is  mounted  on  wheels  so  that  it  can 
be  run  out  in  the  open  for  repairing  and  cleaning. 


BOILER   ACCESSORIES  AND   AUXILIARIES 


47 


The  coal  is  fed  to  the  front  of  the  grate  from  a  hopper  which 
extends  across  the  entire  width  of  the  grate.  At  the  rear  of  the 
hopper  there  is  a  plate  lined  with  firebrick  that  may  be  raised 
or  lowered,  thus  regulating  the  depth  of  fuel-bed.  The  volatile 
matter  in  the  fuel  is  distilled  off  as  the  coal  first  enters  the  fur- 
nace. These  volatile  products  pass  back  over  the  part  of  the 
fire  where  the  fixed  carbon  is  burning,  and  are  given  a  chance  to 
burn  there.  By  the  time  the  fuel-bed  has  reached  the  rear  of 
the  furnace  the  combustion  should  be  complete.  The  ash  and 
clinker  are  dropped  off  to  the  ash-pit  at  the  rear. 


STirry.     ca-  t  •"'     o.°  :    •'eg'     .-ara>o   — o     ^-^^L 

'6'/l'6     ^0  A  d^ "tffiuLllAJiH      ha<i°A'  ° '  A'M  rc'-'MV^AS^^ 


FIG.  18 

The  front  part  of  the  grate  is  overhung  by  a  firebrick  arch. 
This  allows  sufficient  time  and  temperature  for  complete  com- 
bustion before  the  gases  strike  the  heating  surface  of  the  boiler. 

Incomplete  combustion  is  apt  to  occur  with  a  chain  grate  if 
the  fire  is  forced  very  hard.  Excess  air  is  likely  to  leak  in  if  the 
fuel-bed  becomes  too  thin.  This  causes  a  drop  in  the  tempera- 
ture of  the  combustion  chamber  and  therefore  poor  combustion. 

Under  a  light  load,  the  fuel  is  often  burned  before  it  reaches 
the  rear  of  the  grate.  Air  is  likely  to  leak  through  the  ash, 
causing  poor  economy.  As  a  remedy  for  this  condition,  a  con- 
trivance similar  to  a  damper  is  sometimes  placed  under  the  rear 
portion  of  the  grate.  This  makes  it  possible  to  shut  off  the  air 
supply  from  this  part  of  the  grate. 


48 


ENGINES   AND   BOILERS 


46.  The  Roney  Stoker.  —  The  inclined  grate  stoker  is  one  in 
which  the  coal  is  fed  from  a  hopper  at  the  top,  the  coal  burning 


on  its  way  down  across  the  sloping  grate.  In  the  forms  custom- 
arily used,  the  grates  are  operated  mechanically.  There  are  two 
classes  of  these  grates,  side-feed  and  front-feed. 


BOILER   ACCESSORIES   AND   AUXILIARIES  49 

Figure  19  shows  the  Roney  type  of  front-feed  stoker.  The 
coal  is  fed  into  the  hopper,  usually  by  gravity  from  bins  above. 
A  reciprocating  pusher  forces  the  coal  from  the  hopper  onto  a 
dead  plate  beneath  the  front  of  the  arch,  where  the  distillation 
starts.  From  this  plate,  it  is  made  to  move  downward  by  the 
motion  of  the  grates.  By  the  time  the  fuel  reaches  the  ash  plate 
at  the  bottom  of  the  incline  the  combustion  is  complete.  The 
ash  is  dumped  into  the  pit  below  from  time  to  time  by  means  of 
a  hand-lever  that  is  operated  from  the  front  of  the  furnace.  The 
grates  are  rocked  by  means  of  an  eccentric  placed  on  a  rotating 
shaft  running  horizontally  along  the  front  of  the  whole  battery 
of  boilers.  The  amount  the  grates  are  rocked,  and  the  amount 
of  coal  fed,  are  under  the  control  of  the  fireman. 

Since  the  distilled  products  are  driven  off  at  the  front  of  the 
firebrick  arch,  they  have  time,  and  are  at  such  a  temperature, 
due  to  the  fire  below,  that  complete  combustion  takes  place.  In 
some  cases,  steam  and  air  are  admitted  under  the  front  of  the 
arch  to  aid  in  the  combustion.  The  length  of  arch  varies  with 
the  grade  of  fuel  to  be  used,  and  with  the  kind  of  boiler  under 
which  the  stoker  is  installed.  In  some  cases  it  covers  the  entire 
grate,  forming  a  Dutch  oven  which  sits  out  in  front  of  the  rest 
of  the  boiler  setting. 

47.  The  Underfed  Furnace.  —  In  the  types  of  stokers  previ- 
ously described,  the  volatile  matter  is  distilled  and  burnt  over 
the  bed  of  burning  fixed  carbon.  As  the  feeding  of  fuel  is  uniform 
the  amount  of  gas  given  off  at  any  one  time  is  not  so  great  as  in 
hand  firing.  Hence  combustion  has  a  chance  at  all  times  to  be 
more  nearly  complete.  Another  and  radically  different  method 
is  to  feed  the  green  coal  from  beneath,  blowing  the  volatilized 
matter  along  with  sufficient  air  up  through  the  hot  fuel-bed  above, 
where  complete  combustion  takes  place. 

Figure  20  shows  such  a  stoker.  The  coal  is  fed  into  a  hopper 
and  is  forced  back  under  the  fuel-bed  into  retorts,  by  means  of  a 
ram  or  plunger.  Air  under  pressure  is  forced  in  through  tuyeres 
at  the  distillation  zone,  and  by  the  time  the  gases  pass  through 
the  hot  fire  of  fixed  carbon  and  reach  the  top  of  the  fire,  they 
are  completely  burned.  This  type  requires  a  forced  draft.  The 
refuse  is  forced  back  and  down  onto  the  dump  plates  at  the  rear 
of  the  wind-box,  and  is  dropped  from  there  into  the  ash-pit  below 
through  an  adjustable  opening. 


50 


ENGINES  AND   BOILERS 


BOILER   ACCESSORIES  AND   AUXILIARIES  51 

Since  the  temperature  of  the  fire  is  very  high,  most  of  the  ash 
is  fused.  In  order  to  prevent  clogging  it  is  sometimes  necessary 
to  have  a  water-cooled  bridge  wall. 

Due  to  the  forced  draft,  a  plant  thus  equipped  is  not  subject 
to  the  variations  due  to  changing  weather  conditions.  The  rate 
of  combustion  is  regulated  by  the  amount  of  coal  fed  and  the 
amount  of  air  blown  in,  which  are  controlled  together.  The  under- 
feed type  of  stoker  therefore  is  more  flexible  and  will  give  higher 
efficiency  under  conditions  of  forced  load  than  those  previously 
described. 

48.  Smoke  Prevention.  —  Soft  coal  is  generally  considered  to 
be  smoky.  Nevertheless,  it  is  possible  to  burn  practically  all 
grades  of  bituminous  coal  with  very  little  smoke.  Only  during 
the  past  few  years  has  there  been  any  very  determined  effort  to 
rid  ourselves  of  the  smoke  nuisance.  Several  of  our  larger  cities 
have  ordinances  which  are  enforced  more  or  less  rigidly  against 
the  excessive  emission  of  black  smoke  by  power  plants. 

The  soot,  which  is  the  solid  and  black  part  of  the  smoke,  dis- 
figures buildings  and  may  even  injure  health  by  keeping  out  the 
sunlight  and  by  clogging  up  the  respiratory  organs.  It  is  fre- 
quently stated  that  a  great  amount  of  fuel  goes  to  waste  in  the 
soot  of  the  smoke.  While  the  soot  may  be  a  nuisance,  yet  it 
represents  in  itself  but  little  heat  loss.  Soot  indicates,  however, 
incomplete  combustion,  which  often  means  that  there  is  a  large 
amount  of  unburned  hydrocarbon  and  probably  some  CO  that 
should  be  burned  to  C02-  Smoke  prevention,  in  many  cases, 
results  in  an  increased  economy  of  the  power  plant.  Thus  we 
often  find  instances  of  plants  that  have  installed  modern  equip- 
ment which  prevents  the  formation  of  smoke,  not  so  much  with 
the  idea  of  eliminating  the  smoke  as  to  obtain  better  efficiency 
by  means  of  proper  combustion. 

Let  us  consider  in  a  general  way  the  causes  of  smoke  produc- 
tion. Bituminous  coal,  upon  being  heated  to  moderate  tempera- 
tures (600°  to  1000°  F.),  will  have  certain  hydrocarbons  driven 
from  it  in  a  volatile  state.  This  volatile  matter  will  burn  when 
mixed  with  a  sufficient  amount  of  air  and  raised  to  a  sufficiently 
high  temperature  (1800°  to  2000°  F.),  forming  carbon  dioxide 
and  water.  If  for  any  reason  there  is  an  insufficient  supply  of 
oxygen,  or  if  the  hydrocarbons  are  not  raised  to  a  sufficiently  high 
temperature,  incomplete  combustion  will  follow,  and  black  smoke 


52  ENGINES  AND   BOILERS 

may  result.  The  fixed  carbon  left  in  the  coal  after  the  volatile 
hydrocarbons  have  been  driven  off,  with  the  addition  of  sufficient 
air,  will  burn  to  carbon  dioxide.  If  there  is  a  lack  of  air,  carbon 
monoxide,  or  a  mixture  of  carbon  monoxide  and  carbon  dioxide, 
will  result. 

In  conclusion,  to  burn  bituminous  coal  smokelessly,  a  furnace 
must  have  a  sufficient  supply  of  air  to  insure  the  complete  com- 
bustion of  the  volatile  matter,  and  it  must  have  a  temperature 
high  enough  to  permit  of  this  combustion.  In  the  ordinary  fur- 
nace the  time  taken  for  the  gases  to  pass  from  the  grate  to  the 
comparatively  cool  heating  surface  of  the  boiler,  where  they  are 
rapidly  cooled,  is  quite  small,  perhaps  less  than  a  second.  At 
least,  not  enough  time  is  allowed  for  the  volatile  matter  to  unite 
properly  with  the  oxygen  of  the  air,  and  black  smoke  is  the  result. 
If  the  path  of  the  hot  gases  can  be  made  longer,  thus  giving  them 
time  to  burn,  a  large  reduction  can  be  made  in  the  amount  of  smoke. 
It  is  sometimes  possible  to  rearrange  the  baffling  in  the  path  of 
the  gases  in  such  a  way  that  this  is  accomplished  easily.  Another 
way  to  lengthen  the  time  of  burning  is  to  move  the  grates  out 
from  under  the  boiler  and  place  a  long  firebrick  arch  over  the  fire. 

In  ordinary  hand  firing,  a  large  quantity  of  cold  coal  is  thrown 
on  top  of  the  fire,  with  the  result  that  the  fire  is  greatly  cooled, 
both  to  heat  up  the  coal  and  also  to  cause  distillation  of  the 
volatile  matter.  This  reduces  the  temperature  to  a  point  at 
which  the  complete  combustion  of  the  volatile  matter  will  not 
take  place,  and  smoke  results.  At  the  time  the  volatile  matter 
is  given  off  an  excess  of  air  is  needed  to  burn  it.  If  this  air  is 
let  in  over  the  top  of  the  fire  it  will  still  further  cool  the  fire, 
which  only  adds  to  the  trouble. 

The  use  of  the  various  over-feed  stokers,  such  as  the  chain 
grate,  and  those  with  the  front  or  side  feed,  is  a  decided  improve- 
ment over  hand  firing  because  the  fresh  coal  is  added  continu- 
ously, and  the  air  supply  can  be  properly  adjusted  and  main- 
tained. In  these  types  of  stokers,  the  fresh  coal,  upon  coming 
to  the  furnace,  passes  through  the  distillation  period  in  such  a 
position  that  the  volatile  products  must  pass  over  the  fire  of  fixed 
carbon  and  be  burnt  there. 

In  the  underfeed  type  of  stoker,  the  fresh  coal  is  forced  in  from 
the  under  side  of  the  fire,  and  the  distilled  products  along  with 
sufficient  air  to  burn  them,  are  forced  to  pass  through  the  hot 


BOILER   ACCESSORIES   AND    AUXILIARIES  53 

bed  of  burning  carbon  above,  with  the  result  that  a  sufficiently 
high  temperature  is  maintained  to  allow  for  their  complete  com- 
bustion. 

The  purpose  of  the  down-draft  furnace  is  much  the  same  as 
that  of  the  underfeed  stoker.  In  this  type,  the  zones  of  distill- 
ation and  of  burning  the  fixed  carbon  are  separated.  The  first 
takes  place  on  the  upper  grate  and  the  second  on  the  lower;  the 
distilled  product  passes  over  the  hot  fuel  bed  on  the  lower  grate 
and  complete  combustion  occurs. 

Many  devices  to  prevent  smoke  are  on  the  market.  Most  of 
them  consist  of  some  form  of  steam  jet  that  carries  in  and  mixes 
a  sufficient  amount  of  air  with  the  volatilized  hydrocarbons  to 
effect  complete  combustion.  The  steam  itself  has  no  power  to 
prevent  smoke.  It  is  used  simply  to  carry  the  air  and  to  mix 
it  with  the  volatile  products.  If  the  steam  jet  is  left  on  too  long 
after  the  period  of  distillation,  a  loss  greater  than  the  gain 
effected  by  the  jets  may  result.  Some  makers  use  a  dash-pot  or 
other  arrangement  that  automatically  shuts  the  steam  off  soon 
after  each  firing. 

49.  Settings.  —  The  brickwork  that  surrounds  a  boiler  is  called 
a  setting.  The  outer  side  of  this  setting  is  built  of  common  red 
brick.  The  inner  surfaces  that  are  exposed  to  the  high  tempera- 
ture of  the  flame  are  lined  with  firebrick.  With  the  very  high 
temperatures  that  exist  in  modern  furnaces,  it  is  difficult  to  get 
a  grade  of  firebrick  that  will  give  satisfactory  service.  It  is  better 
practice  not  to  leave  an  air  space  between  the  outer  wall  and  the 
lining,  since  heat  is  transmitted  through  the  air  space,  under  the 
high  temperatures  that  exist  in  a  furnace,  faster  than  it  would 
be  through  the  same  thickness  of  brick. 

In  most  furnaces,  a  firebrick  arch  is  placed  over  the  fire.  This 
arch  forms  a  chamber  in  which  the  temperature  is  kept  very  high. 
The  length  of  this  arch  varies  with  the  kind  of  fuel  to  be  used 
and  with  the  type  of  boiler.  Firebrick  baffles  are  placed  between 
the  tubes  of  water  tube  boilers  in  such  a  manner  that  the  gases 
are  forced  to  pass  the  heating  surface  several  times  on  their  way 
to  the  stack.  The  burnt  gases  pass  from  the  setting  to  the  stack 
through  a  duct  called  the  breeching. 

Since  there  is  a  difference  in  pressure  between  the  outside  and 
inside  of  the  setting,  it  is  important  that  there  be  no  cracks  for 
the  air  to  leak  in  or  for  the  gases  to  leak  out.  Under  natural 


54  ENGINES   AND    BOILERS 

draft,  there  is  a  leakage  of  air  inward,  which  cools  the  boiler 
and  injures  the  draft.  With  a  forced  draft,  the  gases  may  leak 
out  into  the  boiler  room. 

50.  Draft.  —  Natural  draft  is  obtained  by  means  of  a  stack  or 
chimney.  The  gases  as  they  leave  the  boiler  are  at  a  tempera- 
ture of  from  400°  to  600°  F.  At  this  temperature  they  are  much 
lighter  than  the  outside  air.  Since  the  column  of  gas  in  the  stack 
is  lighter,  it  is  forced  up  from  the  bottom  by  the  heavier  air 
outside.  The  stack  should  be  large  enough  so  that  but  little 
draft  will  be  lost  by  the  friction  between  the  gas  and  the  stack. 
It  should  be  insulated  so  that  little  heat  is  lost  through  the  walls 
of  the  stack. 

There  are  three  kinds  of  stacks  in  use,  steel,  brick,  and  con- 
crete. The  steel  stack  is  cheaper  and  lighter,  but  it  is  expensive 
in  its  upkeep,  since  it  must  be  painted  often  to  prevent  the 
corroding  of  the  plates.  In  the  better  grades  of  steel  stacks,  a 
firebrick  lining  is  used  at  least  a  part  of  the  way  up  to  prevent 
conduction  of  heat  to  the  outside.  Brick  stacks  are  sometimes 
built  of  hard  common  brick,  but  of  late  years  more  often  of 
special  radial  brick.  They  are  sometimes  lined  with  firebrick, 
as  in  the  steel  stack.  Reinforced  concrete  stacks  have  come  into 
use  during  the  past  few  years.  When  properly  put  up,  they  give 
good  service. 

Brick  and  concrete  stacks  must  be  heavy  enough  to  resist  the 
overturning  effort  of  the  wind.  Steel  stacks  are  either  anchored 
and  designed  to  withstand  the  bending  action  due  to  the  wind, 
or  else  they  are  supported  by  guys. 

Where  forced  draft  is  used,  the  stack  need  be  only  high  enough 
to  discharge  its  smoke  above  the  surrounding  buildings.  Forced 
draft  is  obtained  by  means  of  fans  or  blowers  which  force  the 
air  into  the  ash-pit  or  wind-box  and  thence  through  the  fire. 
In  locomotives,  the  forced  draft  is  obtained  by  means  of  nozzles 
through  which  the  exhaust  steam  from  the  engines  is  discharged 
into  the  chimney.  A  draft  caused  by  this  method  is  sometimes 
called  induced  draft. 

The  amount  of  draft  is  measured  in  inches  of  water.  Under 
natural  draft  it  will  increase  in  going  from  the  ash-pit  to  the  stack 
and  at  the  base  of  the  stack  it  will  be  from  0.5  to  1.5  inches. 
Under  forced  draft  the  pressure  in  the  ash-pit  is  greatest,  and  will 
vary  from  1  to  5  inches. 


BOILER   ACCESSORIES   AND    AUXILIARIES  55 

51.  Dampers.  —  A  damper  should  be  interposed  between  each 
furnace  and  the  stack.     The  efficient  operation  of  the  furnace 
necessitates  careful  attention  to  the  damper.     Automatic  damper 
regulators  are  in  use,  but  for  the  best  results  they  should  be  sup- 
plemented by  intelligent  manual  control. 

52.  Safety  Devices.  —  There   are  in  general  three  causes  of 
explosions  of  properly  designed  boilers:    a  weakened  part,  high 
pressure,  and  low  water.     The  first  is  due  to  the  corroding  or 
wearing  away  of  some  part  of  the  structure,  or  to  a  local  over- 
heating due  to  an  accumulation  of  sediment  or  scale.     It  may 
be  due  to  an  undetected  flaw  in  the  materials  entering  into  the 
makeup  of  the  boiler,  or  it  may  be  due  to  carelessness  or  poor 
workmanship  during  construction. 

The  second  cause,  high  pressure,  is  due  to  a  pressure  much  in 
excess  of  that  for  which  the  boiler  was  designed.  This  may  be 
due  to  a  faulty  safety  valve,  or  to  the  ignorance  of  a  fireman. 

The  third  cause,  low  water,  allows  some  of  the  parts  to  get 
overheated  and  therefore  much  weakened.  It  may  exist  unknown 
to  the  fireman  on  account  of  foaming,  which  is  liable  to  cause  an 
untrue  indication  of  the  water  level  in  the  gage  glass,  or  on  ac- 
count of  some  stoppage  in  the  connection  to  the  glass. 

To  safeguard  against  accidents  due  to  a  weak  part,  it  is  neces- 
sary to  have  a  thorough  inspection  both  of  materials  entering  into 
construction  of  the  boiler  and  of  the  workmanship  during  con- 
struction. There  also  should  be  frequent  inspection  after  the 
boiler  is  put  into  service.  The  common  test  for  strength  is  hydro- 
static. Before  being  put  into  service  a  boiler  should  have  water 
pumped  into  it,  and  a  pressure  should  be  reached  much  in  excess 
of  the  working  pressure. 

53.  The  Pressure  Gage.  —  The  pressure  that  exists  in  a  boiler 
is  measured  by  a  steam  gage.     In  this  country,  the  dial  of  the 
steam  gage  is  graduated  to  read  in  pounds  per  square  inch.     The 
gage  almost  universally  used  is  known  as  the  Bourdon  gage. 
Figure  21  shows  its  internal  construction.     The  pressure  is  ad- 
mitted to  a  curved  flattened  tube  which  is  closed  at  its  free  end. 
This  internal  pressure  tends  to  make  the  curved  tube  straighten 
out.     The  free  end  is  connected  to  the  needle  by  means  of  levers, 
a  rack,  and  a  pinion.     Any  movement  of  the  free  end  causes  the 
hand  or  needle  to  turn,  and  the  pressure  causing  the  movement 
is  indicated  on  the  properly  graduated  dial.     The  flattened  tube 


56 


ENGINES   AND   BOILERS 


is  made  of  brass  or  steel.  Since  a  change  in  the  temperature  of 
the  tube  would  cause  an  error  in  the  reading  of  the  gage,  it  should 
be  connected  to  the  boiler  or  steam-pipe  by  means  of  a  siphon 
so  that  steam  may  never  enter  the  gage.  On  locomotives,  where 
the  gage  is  subject  to  continual  and  severe  jarring,  two  stiff er 
tubes  are  used  in  place  of  the  one  shown  in  the  figure.  The  better 
grades  of  gages  have  a  light  hairspring  to  take  up  the  backlash 
in  the  levers  and  gears. 

A  gage  similar  to  the  one  shown  in  Fig.  21  is  often  used  to 
indicate  vacuum.     For  a  vacuum,  the  tube  is  bent  still  more  and 

the  levers  are  so  arranged  that 
the  motion  of  the  needle  is  re- 
versed from  that  of  the  one  shown 
in  Fig.  21.  The  dials  of  vacuum 
gages  are  commonly  graduated  to 
read  in  inches  of  mercury.  Where 
pressures  are  had  that  may  fluctu- 
ate from  a  vacuum  to  a  positive 
pressure,  gages  are  used  that  will 
indicate  either  the  vacuum  or  the 
pressure  above  the  atmosphere. 
Gages  should  be  tested  from  time 
to  time  to  see  that  they  give  the 
correct  pressure  reading. 
54.  The  Safety  Valve.  —  The  purpose  of  a  safety  valve  on  a 
boiler  is  to  prevent  an  undue  or  dangerous  pressure.  It  is  im- 
possible in  an  ordinary  furnace  to  regulate  the  combustion  quickly 
enough  to  correspond  to  sudden  changes  of  the  amount  of  steam 
used.  For  instance,  a  boiler  may  be  furnishing  its  maximum 
amount  of  steam,  when  for  some  reason  the  engine  is  shut  down 
without  warning,  and  therefore  before  the  fire  can  be  deadened. 
As  a  result,  the  rapid  formation  of  steam  will  continue  long 
enough  to  cause  an  excessive  boiler  pressure  if  the  safety  valve 
does  not  give  relief.  Hence  the  safety  valve  should  have  such  a 
capacity  that  it  is  capable  of  discharging  all  the  steam  that  the 
boiler  can  generate  without  allowing  the  pressure  to  become 
dangerous.  Furthermore  the  safety  valve  must  be  absolutely 
reliable  in  its  action,  and  it  should  be  so  constructed  and  placed 
on  the  boiler  that  it  cannot  be  put  out  of  action  through  careless- 
ness or  ignorance.  No  stop  valve  should  be  placed  between  it 


FIG.  21 


BOILER   ACCESSORIES   AND   AUXILIARIES 


57 


and  the  boiler.  Many  explosions  have  been  caused  by  the  failure 
of  the  safety  valve  to  operate.  As  far  as  the  writer  knows,  how- 
ever, none  have  occurred  that  were  due  entirely  to  excessive 
pressure  when  the  valve  was  in  action.  Several  types  of  safety 
valve  have  been  used  in  the  past,  but  with  the  pressure  ordi- 
narily carried  in  this  country,  the  use  of  the  pop  type  has  become 
almost  universal,  and  will  be  the  only  one  described  here. 

Figure  22  shows  in  section  a  pop  safety  valve.  Most  safety  valves 
are  made  with  a  45°  seat.  The  valve  is  held  on  the  seat  by  a 
helical  spring.  When  the  steam  pressure  becomes  sufficient  to 
overcome  the  force  of  the  spring,  the  valve  is  raised  enough  to  al- 
low some  steam  to  escape.  This  steam  passes  into  a  huddling 
chamber.  The  area  upon  which  the  steam  now  acts  is  slightly 
greater  than  before  the  valve  opened,  with  the  result  that  the  spring 
is  compressed  suddenly  to  a  greater  extent  than  if  the  steam  acted 
only  upon  the  original  area.  Escape  of  steam  will  continue  until 
the  pressure  has  dropped  to  a  few  pounds  less  than  that  at  which 
it  opened.  When  the  valve  stops  blowing,  it  seats  firmly.  Since 
the  pressure  is  less  than  that  at  which  it  opened,  it  will  remain 
shut  until  the  steam  pressure 
again  reaches  the  popping  point. 
The  valve  should  be  constructed 
and  set  so  that  the  difference  be- 
tween the  popping  pressure  and 
the  closing  pressure  or  Slowdown 
is  not  too  large,  in  order  to  pre- 
vent shock  to  the  boiler  and  an 
excessive  loss  of  steam  during 
ordinary  operation. 

A  lever  is  provided  at  the  top 
of  the  casing  for  locking  the 
valve  open.  The  compression  in 
the  spring  may  be  adjusted  by 
screwing  the  top  cap  up  or  down. 
In  most  valves  the  spring  is  en- 
cased so  as  to  protect  it  from 
the  escaping  steam,  and  to  pre- 
vent back  pressure  in  the  dis- 
charge pipe  from  acting  on  the 
top  of  the  valve.  (See  Fig.  22.)  FIG.  22 


58  ENGINES   AND   BOILERS 

55.  Safety-valve  Capacity.  —  The  amount  of  steam  that  a 
safety  valve  discharges  depends  upon  the  steam  pressure  and  upon 
the  effective  opening.  The  latter  varies  with  the  lift  of  the  valve. 
Not  all  valves  of  the  same  diameter  have  the  same  lift;  hence 
they  differ  in  capacity.  There  has  been  considerable  agitation 
recently  to  have  all  pop  valves  put  upon  a  uniform  rating.  Tests 
have  been  made  to  determine  the  discharge  and  the  lift  under 
various  conditions  of  pressure.  These  tests  show  that  the  com- 
monly used  empirical  formula  given  by  Napier  is  substantially 
correct  when  applied  to  the  safety  valve. 


56.  Napier's  Formula.  —  Napier's  formula  for  steam  issuing 
from  an  orifice  into  the  atmosphere  is 

W-A'P 

"W 

in  which  W  is  the  weight  in  pounds  of  steam  issuing  per  second, 
A  is  the  area  of  the  orifice  in  square  inches,  and  P  is  the  abso- 
lute pressure  in  pounds  per  square  inch  in  front  of  the  orifice. 

Applying  this  formula  to  the  safety  valve  with  a  45°  seat,  it 
is  seen  that  the  area  oi  opening,  A,  equals  approximately  the 
product  of  irD  and  the  lift  times  the  sine  of  45°,  where  D  is  the 
diameter  of  the  valve  in  inches.  If  the  lift  is  known,  the  dis- 
charge may  be  calculated.  Some  valve  makers  use  the  assump- 
tion that  the  lift  is  1/30  of  the  diameter,  and  rate  their  valves 
accordingly.  Under  this  assumption  it  is  seen  that 

7rD2X.707xP 


W  = 


70X30 


and  the  weight  of  steam  discharged  per  hour  is  3600  W  =  3.81  PD2. 
As  previously  explained,  the  safety  valve  must  be  large  enough 
to  discharge  the  maximum  amount  of  steam  the  boiler  is  capable 
of  generating.  We  can  compute  this  maximum  amount  from  the 
heating  surface  of  the  boiler,  allowing  an  evaporation  of  from  six 
to  ten  pounds  of  water  per  square  foot  of  heating  surface  per 
hour,  or  we  may  compute  it  from  the  grate  area,  assuming  a 
boiler  efficiency  and  a  rate  of  combustion  consistent  with  the 
draft  and  with  the  kind  of  fuel  used.  The  former  method  is  con- 
sidered better.  After  conducting  numerous  tests  P.  G.  DARLING1 

i  TRANS.  A.  S.  M.  E.,  vol.  31  (1909),  p.  109. 


BOILER   ACCESSORIES   AND   AUXILIARIES  59 

advocated  to  the  A.  S.  M.  E.  formulas  for  pop  safety  valves 
derived  according  to  the  following  method: 

TT 

for  stationary  boilers,  D=  0.068  ^j 

-L/Z 

TT 

for  locomotive  boilers,  D=  0.055  f^> 

Lir 

in  which  D  is  the  diameter  of  the  valve  in  inches,  H  is  the  heat- 
ing surface  of  the  boiler  in  square  feet,  L  is  the  lift  of  the  valve 
in  inches,  and  P  is  the  absolute  boiler  pressure  in  pounds  per 
square  inch.  It  is  noticed  that  smaller  valves  are  required  for 
locomotives,  because  the  maximum  draft  can  be  secured  only 
when  the  steam  is  being  drawn  from  the  boiler  by  the  engine. 

57.  Other  Safety-valve  Formulas.  —  Various  cities  and  states 
have  their  own  rules  governing  the  sizes  of  safety  valves,  a  few 
of  which  are  given  below. 

CITY  OF  CHICAGO.  One  square  inch  of  pop-valve  area  (7rD2/4) 
for  every  three  square  feet  of  boiler  grate  area. 

CITY  OF  PHILADELPHIA.  For  pop  valves,  A  =  22.5XCr/(p  —  8.62), 
in  which  A  is  the  area  of  the  valve  in  square  inches  (not  the  effec- 
tive opening  for  the  escape  of  steam),  G,  the  grate  area  in  square 
feet,  and  p  the  gage  boiler  pressure. 

U.  S.  SUPERVISING  INSPECTORS.  A  =  .2Q74:XWH/P,  in  which  A 
is  the  area  of  the  valve  as  in  the  previous  formula,  WH  is  the 
number  of  pounds  of  water  evaporated  by  the  boiler  per  hour, 
and  P  is  the  absolute  boiler  pressure. 

Safety  valves  are  not  made  in  sizes  over  5  or  6  inches  in  diameter. 
In  large  boilers  it  is  therefore  necessary  to  use  more  than  one. 

EXAMPLE.  What  should  be  the  size  of  pop  safety  valve  on  a  boiler  with 
1500  square  feet  heating  surface  if  the  pressure  carried  is  130  pounds  gage? 

SOLUTION.  Assuming  a  maximum  rate  of  evaporation  of  8  pounds  of 
water  per  square  foot  of  heating  surface  per  hour,  we  get  8X1500  =  12000 
pounds  of  steam  to  be  discharged  per  hour  through  the  valve.  The  weight 
discharged  per  second  is  12000/3600  =  3.33  pounds.  From  Napier's  formula, 
W  =  AP/70,  we  see  that  3.33  =AX(130+15)/70,  whence  A,  the  area  of 
opening,  in  square  inches,  is  1.61  Assuming  a  lift  of  valve  equal  to  1/30  of 
the  diameter,  the  area  of  the  opening  will  be  approximately  .7077rD2/30. 
Then  .7077rZ)2/30  =  1.61,  or  D  =  4.67  inches.  Hence  a  5-inch  valve  should  be 
used. 

58.  The  Water  Glass  or  Gage  Glass.  —  In  order  that  the 
amount  of  water  in  the  boiler  may  be  known,  a  water  glass  is 
attached.     The  lower  end  of  the  water  glass  is  attached  to  the 


60 


ENGINES   AND   BOILERS 


water  space,  and  the  upper  to  the  steam  space.  Since  there  is 
danger  of  the  glass  becoming  stopped  in  these  connections,  and 
the  water  level  thereby  being  falsely  indicated,  or  of  the  glass 

being  broken,  three  gage-cocks 
or  try-cocks  are  placed  on  the 
boiler  or  water  column.  The 
top  cock  is  placed  above,  and 
the  bottom  one  below,  the 
normal  water  level.  By  open- 
ing these  cocks  in  succession, 
one  may  determine  whether 
or  not  the  gage  glass  is  giving 
the  correct  level. 

59.  High  and  Low  Water 
Alarm.  —  High- water  and  low- 
water  alarms  are  sometimes 
used  to  attract  the  attention 
of  the  fireman  when  the  water 
falls  below  or  rises  above  the 
safe  level.  The  alarm  is 
operated  by  means  of  a  float 
in  the  water  column.  When 
this  float  rises  too  high  or  falls 
too  low  it  will  open  a  valve, 
and  allow  the  escaping  steam  to  blow  a  whistle.  (See  Fig.  23.) 

60.  Fusible  Plug.  —  Another  safety  device  used  to  detect  low 
water  is  the  fusible  plug.  This  plug  (Fig.  24)  has  a  tin  core 
that  will  melt  when  the  water 
level  falls  below  it.  These 
plugs  are  placed  in  the  crown 
sheet  of  an  internally  fired 
boiler  in  the  rear  head  a  little 
above  the  top  tubes  in  the 
return-tubular  boiler  and  in  the  bottom  of  the  steam  drum  of  a 
water-tube  boiler.  They  should  be  kept  free  from  scale  on  the 
inside  and  from  soot  on  the  outside.  None  of  the  above  safety 
devices  are  absolutely  certain  in  their  action.  Under  conditions 
of  very  rapid  steaming  or  with  feedwater  that  foams,  the  water 
level  in  the  boiler  may  be  below  that  in  the  water  column. 


FIG.  23 


//7S/t/f  or  arfsjure 
iat/Se  ry/*e 


Outft'efe  ft/pes 
0ttts/e/e  or  // 

FIG.  24 


BOILER   ACCESSORIES   AND   AUXILIARIES  61 

61.  Boiler  Feedwater  Treatment.  —  The  impurities  in  water 
that  are  responsible  for  most  of  the  scale  formation  are  the  car- 
bonates and  sulphates  of  calcium  and  magnesium.     If  muddy 
water  is  used,  the  mud  may  be  deposited  on  the  heating  surface 
and  aid  in  scale  formation. 

The  carbonates  of  lime  and  magnesium  are  soluble  in  water 
containing  carbon  dioxide.  These  carbonates  cause  what  is  known 
as  temporary  hardness. .  Upon  heating  to  212°  F.  the  carbon 
dioxide  is  driven  off,  and  the  carbonates  are  precipitated.  If 
these  are  the  only  impurities  in  the  feedwater,  an  open  feedwater 
heater  will  remove  most  of  the  scale-forming  material.  Where  a 
heater  is  not  used  the  carbonates  may  be  precipitated  by  the 
addition  of  a  solution  of  slacked  lime.  The  lime  combines  with 
the  carbon  dioxide  to  form  the  insoluble  monocarbonate  of  lime. 

The  sulphates  of  lime  and  magnesium  are  not  precipitated  at  a 
temperature  of  212°,  but  are  precipitated  at  a  temperature  such 
as  exists  in  the  boiler.  They  cause  what  is  known  as  permanent 
hardness.  The  addition  of  carbonate  or  hydrate  of  soda  (or  a 
mixture  of  the  two)  will  cause  precipitation.  The  carbonate  of 
soda  decomposes  the  sulphates  and  forms  insoluble  carbonates  of 
lime  and  magnesium,  which  precipitate,  leaving  neutral  soda  and 
sodium  sulphate  in  solution.  If  carbon  dioxide  is  present,  the 
soluble  bicarbonate  of  lime  is  formed,  which  may  be  precipitated 
by  heating  or  by  the  addition  of  lime  as  explained  previously.  In 
most  purification  processes  both  the  lime  and  soda  are  used. 

If  organic  matter,  from  sewage  or  from  some  other  source,  is 
present  in  the  water,  it  may  be  removed  by  filtration.  Before 
passing  the  filter  a  coagulant,  such  as  alum,  is  often  used.  Organic 
matter  in  feedwater  is  often  the  cause  of  foaming. 

62.  Scale  Prevention  and  Removal.  —  Many  substances  have 
been  used  to  prevent  the  formation  of  scale.      Some  of  these 
probably  do  as  much  damage  to  the  boiler  as  would  the  scale. 
Aside  from  the  treatment  to  remove  the  scale-forming  material, 
the  best  substance  seems  to  be  graphite.     When  it  is  injected 
into  the  boiler,  it  is  said  to  help  in  the  prevention  of  scale  formation. 

Where  no  precaution  is  taken  to  prevent  the  scale  from  form- 
ing, it  is  necessary  to  clean  it  from  the  tubes  periodically.  This 
is  usually  done  by  means  of  a  cutter  or  hammer  that  is  driven 
by  a  small  air,  steam,  or  water  turbine. 


62 


ENGINES   AND   BOILERS 


Figure  25  shows  one  make  of  cleaner  that  is  applied  to  a  fire 
tube.     Figure  26  shows  the  form  that  is  applied  to  a  water  tube. 


FIG.  25 


Jca/a 


tfrr/zeafaf 
jecf/o/J 


t/rfer  fi/bc 

FIG.  26 

63.  Oil  Separators.  —  In  plants  where  all  or  part  of  the  steam 
is  condensed  and  used  again  as  boiler  feed,  the  oil  that  was  used 
to  lubricate  the  engine  will  find  its  way  to  the  boiler.  This  does 
not  apply  to  steam  turbines,  as  oil  is  not  usually  used  internally 
with  them.  This  oil  forms  a  very  hard  scale  that  it  is  almost  im- 

possible  to  remove.  To  pre- 
vent this,  oil  separators  are 
used  to  remove  the  oil,  either 
from  the  exhaust  steam  or 
from  the  water  after  it  is  con- 
densed. The  removal  before 
condensation  is  preferable, 
since  the  oil  does  not  have  to 
be  contended  with  in  the  con- 
denser or  feedwater  heater. 
Figure  27  shows  an  Austin  oil 
separator.  Since  the  separator 
is  quite  large,  the  steam  passes 
through  it  with  a  small  veloc- 
ity and  deposits  the  oil  on  the 
surface  of  the  corrugated  ver- 
tical baffle  plate  shown  in  plan 
and  section  in  the  figure. 
With  high  vacua,  a  spray  of 
water  keeps  the  surface  moist, 
which  aids  in  the  separation 
FIG.  27  of  the  oil. 


BOILER   ACCESSORIES   AND   AUXILIARIES 


63 


64.  Boiler  Feed-pumps.  —  The  feedwater  is  usually  forced  into 
a  boiler  by  means  of  a  pump.  Figure  28  shows  a  common  type 
of  boiler  feed-pump.  This  pump  is  direct  acting,  the  steam  piston 
and  water  piston  or  plunger  being  fastened  to  the  same  piston  rod. 
Since  the  steam  and  the  water  in  a  boiler  are  both  under  the 
same  pressure  and  since  pipes  and  fittings  offer  a  resistance  to 
the  flow  of  both  water  and  steam,  it  is  seen  that  it  is  necessary  to 


FIG.  28 


make  the  steam  piston  of  the  feed-pump  considerably  larger  in 
diameter  than  the  water  plunger. 

Steam  is  admitted  by  the  steam  valve  alternately  to  the  two 
ends  of  the  steam  cylinder.  At  the  same  time  that  the  valve 
is  admitting  steam  to  one  end  of  the  cylinder  it  is  allowing  it  to 
exhaust  from  the  other,  thus  giving  the  piston  a  reciprocating 
motion.  The  water  piston  sucks  up  water  from  the  suction  pipe 
in  one  end  of  the  water  cylinder  while  forcing  it  into  the  delivery 
pipe  on  the  other. 

The  suction  pipe  leads  either  to  the  hot  well  or  to  the  cold  well 
from  which  the  feedwater  is  taken.  The  lower  end  of  the  suction 
pipe  should  be  provided  with  a  strainer  to  prevent  any  large 
pieces  of  solid  matter  from  getting  into  and  clogging  the  valves. 
A  foot  valve  should  be  provided  also  to  keep  the  suction  pipe 


64  ENGINES   AND   BOILERS 

full  of  water  when  the  pump  is  not  running,  thus  eliminating  the 
priming  of  the  pump  every  time  it  is  started. 

There  are  two  sets  of  valves  at  each  end  of  the  water  cylinder. 
On  the  suction  stroke,  the  suction  valves  are  lifted  and  the  water 
is  sucked  in  back  of  the  piston.  During  the  forcing  stroke,  these 
valves  are  closed  and  the  water  is  forced  out  through  the  upper 
set  and  into  the  delivery  pipe.  A  light  spring  is  employed  to 
help  seat  the  valve.  Most  valves  are  faced  with  a  composition 
disc  which  may  be  replaced  when  it  becomes  worn. 

The  flow  from  a  reciprocating  pump  is  not  steady;  to  insure 
a  more  uniform  rate  of  flow  of  the  water  an  air  chamber  is  placed 
on  the  delivery  line  close  to  the  pump.  This  is  kept  partly  filled 
with  air,  which  acts  as  a  cushion.  A  check  valve  is  placed  in  the 
feedline  between  the  pump  and  the  boiler.  This  prevents  the 
water  from  the  boiler  escaping  back  through  leaky  valves  when 
the  pump  is  not  in  full  operation.  There  should  also  be  a  stop 
valve  in  the  feedline. 

There  are  two  types  of  reciprocating  steam  pump,  one  in  which 
there  is  only  a  single  steam  cylinder  and  a  single  water  cylinder, 
and  the  other  in  which  there  are  two  steam  cylinders  and  two 
water  cylinders  placed  side  by  side.  The  latter  type  is  called  a 
duplex  pump.  In  this  type  the  valve  for  one  steam  cylinder  is 
operated  by  the  movement  of  the  piston  of  the  other  steam 
cylinder. 

These  boiler  feed-pumps  take  steam  the  full  length  of  the  stroke, 
not  allowing  it  to  expand  in  the  cylinder,  and  they  are  not  eco- 
nomical in  the  use  of  steam.  (See  Chapter  VI  on  the  steam  en- 
gine.) However,  only  a  small  proportion  of  the  steam  generated 
by  the  boiler  is  needed  to  run  the  pump.  To  secure  better 
economy,  feed-pumps  are  occasionally  driven  by  power  taken 
from  the  main  engine.  The  supply  of  water  from  these  pumps 
is  not  easily  regulated.  They  are  made  to  pump  more  water 
than  is  normally  required,  the  excess  being  passed  back  to  the 
suction  through  a  relief  valve.  In  large  electric  power  plants, 
triplex  pumps,  driven  by  electric  motors  are  often  used  for  boiler 
feeding.  Of  late  years  centrifugal  and  turbine  pumps  have  been 
employed  for  boiler  feeding. 

An  automatic  regulator  is  sometimes  installed  with  the  pump 
so  that  the  pump  will  furnish  just  the  proper  amount  of  water 
to  keep  the  boiler  water-level  constant. 


BOILER   ACCESSORIES   AND   AUXILIARIES 


65 


65.  The  Injector.  —  On  portable  boilers  and  in  small  plants, 
the  water  is  often  forced  into  the  boiler  by  means  of  an  injector 
or  inspirator.  This  is  usual  also  on  locomotives.  The  principle 
upon  which  the  injector  works  is  illustrated  by  Fig.  29.  Steam 
from  the  boiler  enters  the  injector  through  a  steam  nozzle,  a,  in 
which  it  expands  and  some  of  its  heat  energy  is  transformed  into 
kinetic  energy.  The  steam  leaves  the  nozzle  with  a  high  velocity 
and  enters  a  small  combining  tube,  b.  The  water  inlet  leads  to  a 
chamber  which  is  located  between  the  nozzle  and  the  combining 
tube.  As  the  steam  flows  from  the  nozzle  to  the  combining  tube 
it  tends  to  form  a  par- 
tial vacuum  in  the 
water  chamber  and 
thus  sucks  up  and  car- 
ries the  water  along 
with  it.  The  steam 
mixes  with  the  water  in 
the  combining  tube  and 
is  condensed.  This  mix- 


ture  of  condensed  steam 
and  water  has  a  high 
velocity  and  therefore 
a  considerable  amount 
of  kinetic  energy;  its 
pressure,  however,  is  " 
atmospheric  or  less. 
This  mixture  passes 
from  the  combining 
tube  to  the  delivery  FIG.  29 

tube,  c,  which  has  an 

increasing  diameter.  The  mixture  therefore  loses  a  large  amount 
of  its  velocity  and  kinetic  energy.  What  it  loses  in  kinetic  energy 
it  gains  in  pressure  energy,  so  that  by  the  time  it  leaves  the  in- 
jector it  has  gained  enough  pressure  to  force  open  the  check 
valve  leading  to  the  boiler.  An  overflow  is  located  at  the  end 
of  the  combining  tube,  so  that  when  steam  is  first  turned  on, 
it  escapes  through  the  overflow.  The  overflow  is  fitted  with  a 
valve  which  automatically  closes  when  the  pressure  inside  the 
combining  tube  falls  below  the  pressure  of  the  atmosphere,  thus  pre- 
venting air  from  coming  into  the  injector  and  impeding  its  action. 


66  ENGINES   AND   BOILERS 

Unless  specially  constructed,  an  injector  cannot  lift  water  a 
very  great  height.  Moreover,  since  the  injector  must  condense 
the  steam  in  order  to  work  at  all,  it  is  necessary  that  the  water 
be  cold.  Considered  as  a  pump,  the  efficiency  of  the  injector  is 
very  low,  because  the  greater  part  of  the  energy  of  the  steam 
goes  to  heat  the  water.  If  it  is  used  to  feed  a  boiler,  the  heat 
spent  in  raising  the  temperature  of  the  feedwater  is  not  lost,  as 
it  goes  back  into  the  boiler.  Hence  it  is  efficient  for  this  purpose. 
The  injector  is  light,  occupies  but  little  space,  and  is  cheaper 
than  a  pump,  but  it  is  not  so  dependable. 

66.  Boiler  Feeding  by  Return  Trap.  —  The  condensation  from 
various  parts  of  the  plant  is  sometimes  returned  to  the  boiler 
by  what  is  known  as  a  return  trap.     This  trap  is  located  above 
the  level  of  the  boiler  and  the  water  runs  into  the  boiler  under 
the  influence  of  gravity  and  the  pressure  of  live  steam.     These 
traps  are  quite  economical  in  the  use  of  steam  and  they  may  be 
used  to  supply  all  the  feedwater.     They  are  not  nearly  so  reliable 
as  the  pump  or  injector,  however,  and  are  therefore  but  little 
used  to  furnish  the  entire  feedwater  supply. 

67.  The  Steam  Line.  —  Steam  pipe  is  made  of  wrought  iron  or 
of  steel.     The  nominal  diameter  corresponds  approximately  with 
the  inside  diameter.     Sizes  of  standard  pipe  vary,  by  the  y%" 
from  y8"  to  y2",  by  the  M"  from  %*  to  IJ^",  by  the  %'  from 
IY2"  to  5",  and  by  the  1"  from  5"  to  15". 

It  has  been  customary  to  allow  an  average  velocity  of  steam 
in  the  line  of  from  4000  to  6000  feet  per  minute.  In  modern 
turbine  plants,  however,  where  the  flow  is  uniform,  and  especially 
where  superheated  steam  is  used,  velocities  much  in  excess  of 
these  values  are  used.  If  a  velocity  of  wet  steam  much  greater 
than  that  just  mentioned  is  used,  the  drop  in  pressure  due  to  skin 
friction  will  be  excessive.  On  the  other  hand,  if  a  velocity  much 
less  is  allowed,  too  large  and  expensive  a  pipe  will  be  required. 

If  the  volume  and  velocity  of  steam  to  be  carried  by  the  pipe 
line  are  known,  the  diameter  is  easily  determined.  The  volume 
carried  per  unit  of  time  equals  the  product  of  the  area  of  the 
cross-section  of  the  pipe  and  the  velocity.  If  the  size  and  speed 
of  the  engine  to  be  supplied  are  known,  we  may  compute  the  vol- 
ume of  steam  needed.  At  maximum  it  may  be  assumed  that  the 
engine  takes  steam  during  the  full  length  of  the  stroke.  When 


BOILER   ACCESSORIES   AND   AUXILIARIES 


67 


more  than  one  boiler  is  used  it  is  customary  to  discharge  the 
steam  into  a  common  pipe  called  a  header.  In  such  a  case 
each  boiler  should  be  provided  with  a  non-return  stop-valve  be- 
tween the  boiler  and  the  header.  This  non-return  stop-valve 
acts  as  a  check  valve  in  case  the  direction  of  steam  flow  should 
be  reversed,  which  would  happen  in  case  a  tube  blew  out  or  some 
other  similar  accident  occurred. 

The  piping  must  be  provided  with  a  sufficient  number  of 
hangers  to  prevent  breaking  due  to  its  own  weight.  The  line 
should  slope  downward  in  the  direction  the  steam  is  to  flow  in 
order  that  the  condensation  may  be  carried  along  with  the  steam. 
If  this  precaution  is  not  followed  condensed  steam  will  collect 
in  the  pipe  and  may  be  carried  in  slugs  by  the  steam  in  amounts 
large  enough  to  injure  and  cause  leakage  or  even  breakage  of 
the  fittings.  Provision  should  be  made  at  the  low  points  of  the 
line  to  remove  condensation.  A  pipe  should  be  run  down  from 
the  low  point  and  the 


Va/re  seaf 
fa/re 


water  collecting  in  this 
may  be  blown  out  from 
time  to  time  by  open- 
ing a  valve  by  hand. 
A  trap  may  be  installed 
that  will  remove  it 
automatically. 

68.  The  Steam  Trap. 

-  Several     types      of 
traps  are  in  use.    In  the 
more    common    kinds, 
the  valve   is  operated 

by  means  of  either  a 

*  T 
float,  the  unequal  ex- 

pansion  of  two  differ- 
ent metals  with  chang- 
ing  temperature,  press- 
ure of  collected  water 
on  a  flexible  diaphragm, 
or  the  weight  of  a  bucket  as  it  fills  with  water.  The  latter  kind 
is  illustrated  in  Fig.  30.  In  this  type  the  buoyancy  of  the 
bucket  keeps  the  valve  closed  until  enough  water  flows  over  the 


FIG.  30 


68 


ENGINES   AND    BOILERS 


edge  and  collects  in  the  bucket  to  sink  it.  The  sinking  of  the 
bucket  opens  the  valve  and  the  water  collected  in  the  bucket  is 
forced  out  through  the  valve  by  the  steam  pressure  inside  the  trap. 
The  bucket  now  being  lightened,  it  again  rises,  closing  the  valve. 

In  many  traps,  the  valve  is  operated  by  a  float.  The  water 
collects  in  a  float  chamber  and  raises  the  buoyant  float  until  the 
valve  is  opened.  The  water  then  escapes  until  the  float  is  lowered 
enough  to  allow  the  valve  to  seat.  An  air  valve  is  located  at  the 
top  of  the  trap  to  allow  the  air  to  escape  if  enough  should  be  caught 
there  to  interfere  with  the  operation  of  the  trap. 

Another  form  of  trap  is  one  in  which  the  valve  is  operated  by 
the  unequal  expansion  of  two  metals.  When  the  trap  is  cold  the 
valve  is  open  and  the  water  is  allowed  to  escape.  As  soon  as 
the  steam  flows  through,  however,  the  parts  are  heated  and  ex- 


FIG.  31 


FIG.  32 


pand  unequally,  closing  the  valve.  Water  then  collects  again, 
and  as  the  parts  cool,  the  valve  will  again  open  and  the  opera- 
tion will  be  repeated.  When  large  amounts  of  water  are  to  be 
handled,  dumping  traps  may  be  used.  The  discharged  water  from 
the  trap  is  led  to  the  drain  or  is  piped  back  to  the  hot  well. 

69.  Expansion  Joints.  —  Since  the  pipe  is  laid  cold,  it  will  ex- 
pand when  steam  is  turned  into  it  and  its  temperature  becomes 
that  of  the  steam.  The  expansion  amounts  to  2.5  inches  per 
hundred  feet  of  pipe  with  ordinary  steam  temperatures,  and  may 
be  greater  when  the  steam  is  of  very  high  pressure  and  is  super- 
heated. The  piping  must  be  so  arranged  that  this  expansion  may 
take  place  without  injury  to  the  pipe.  If  the  pipe  is  not  laid 
straight  but  contains  elbows,  it  may  bend  enough  so  that  no  dan- 
gerous stresses  will  be  induced.  If  there  is  a  considerable  run  of 
straight  pipe,  however,  expansion  joints  must  be  provided. 

There  are  several  types  of  expansion  joints  in  use.  A  very  com- 
mon kind  for  use  with  low-pressure  steam  is  the  slip- joint.  In 


BOILER    ACCESSORIES    AND    AUXILIARIES 


69 


this,  provision  is  made  for  the  slippage  of  one  part  of  the  joint 
on  the  other.  The  joint  is  kept  steam  tight  by  means  of  a  stuf- 
fing box.  Figure  31  shows  this  type.  Goosenecks  and  expan- 
sion loops  (Fig.  32)  are  used  when  the  steam  pressure  is  high. 

70.  Steam  Separators.  —  Unless  superheat  is  used,  steam  leav- 
ing the  boiler  will  always  contain  some  moisture.     If  the  steam- 
pipe  is  very  long,  some  condensation  also  takes  place.     Due  to 
these  causes,  the  steam  is  liable  to  reach  the  engine  quite  wet. 
It  is  desirable  both  for  safety  and  for  economy  to  have  the  steam 
as  dry  as  possible  when  it  enters  the  engine.     To  remove  the 
moisture  from  the  steam,  a  separator  is  placed  in  the  line  just 
before  it  reaches  the  engine.      The  steam  is 

given  a  sudden  change  in  direction  upon  enter- 
ing the  separator.  The  moisture  resists  this 
change  to  a  greater  extent  than  does  the  steam. 
In  the  type  shown  in  Fig.  33  the  steam  is  first 
deflected  downward  and  then  upward,  and  as 
the  moisture  cannot  change  its  direction  of 
motion  as  rapidly  as  the  steam,  it  is  caught 
and  collected  in  the  bottom  of  the  separator. 

In  some  makes  the  steam  is  given  a  whirl- 
ing motion  and  the  water,  being  denser  than 
the  steam,  is  forced  to  the  outside  of  the 
separator,  where  it  is  collected.  Another  type, 
which  is  similar  to  the  oil  separator  of  Fig.  27, 
is  that  in  which  a  corrugated  baffle  plate  is 
interposed  in  the  path  of  the  steam.  The  steam  passes  around 
the  baffle  while  the  moisture  is  caught  by  it  and  runs  down  the 
corrugations  to  the  bottom  of  the  separator,  where  it  is  collected. 

A  separator  should  remove  most  of  the  moisture,  but  it  should 
not  offer  too  great  a  resistance  to  the  passage  of  the  steam,  since 
this  would  cause  a  drop  in  pressure.  The  moisture,  after  being 
collected,  is  trapped  off  and  discharged  to  the  drain  or  returned 
to  the  hot  well.  Often  the  separator  is  made  large  and  acts  as  a 
steam  receiver.  This  reduces  the  pulsation  in  the  steam  line 
when  the  steam  is  used  by  a  reciprocating  engine. 

71.  Steam-pipe  Covering.  —  To  prevent  radiation  of  heat  from 
the  steam-pipe  and  the  consequent  condensation,  a  covering  is 
applied  to  the  pipe.     The  covering  is  made  from  materials  that 


FIG.  33 


70 


ENGINES    AND    BOILERS 


are  poor  conductors  of  heat.  A  finely-divided,  dead  air  space  is 
one  of  the  best  non-conductors  of  heat.  In  most  coverings  the 
object  is  to  get  as  much  finely-divided  dead  air  space  as  possible. 


7  r  <c          y- 


FIG.  34 

The  most  common  kinds  of  covering  are  made  from  asbestos 
or  a  mixture  of  asbestos  and  carbonate  of  magnesia.  The  mag- 
nesia used  in  pipe  covering  contains  a  great  number  of  very  small 
air  cells,  and  therefore  makes  an  excellent  insulator.  When  the 
magnesia  is  used  it  is  usually  moulded  into  hollow  cylinders  to- 
gether with  enough  asbestos  fiber  to  give  it  strength.  Another 
form  of  covering  much  used  is  made  of  several  thicknesses  of 
corrugated  asbestos  paper  formed  into  hollow  cylinders  by  wind- 
ing on  a  mandrel. 


BOILER   ACCESSORIES   AND   AUXILIARIES 


71 


72.  Feedwater  Heaters.  —  In  most  plants,  all  or  some  of  the 
steam  is  exhausted  at  atmospheric  pressure.     If  this  steam  is 
exhausted  to  the  air  all  of  its  heat  is  wasted.     Some  of  this  heat 
may  be  used  to  heat  the  boiler  feedwater  by  running  the  exhaust 
steam  through  a  feedwater  heater  and  extracting  its  heat  of 
vaporization.     There  are  two  types  of 

heater,  the  open  and  the  closed. 

I^the  open  feed^watexjieater  the 
steam  comes  m  direct  contact  with 
the  feedwater,  which  is  made  to  flow 
over  shallow  pans,  thus  exposing  a 
large  area  to  the  steam.  The  tem- 
perature of  the  water  is  thereby 
brought  near  to  the  boiling  point.  If 
the  water  is  hard,  a  large  part  of  the 
scale-forming  materials  will  be  de- 
posited on  the  pans,  which  may  be 
easily  removed  and  cleaned.  Figure 
34  shows  a  common  form  of  open 
heater.  A  skimmer  is  provided  to 
remove  the  oil  that  comes  in  with  the 
exhaust  steam,  and  there  is  also  a  filter 
to  purify  the  feedwater.  The  pur- 
pose of  the  open  heater  is  thus  seen  to 
be  twofold:  to  utilize  the  heat  that 
would  otherwise  be  wasted  and  to 
purify  the  water.  In  the  closed  type 

(Fig.  35),  the  steam  and  the  feedwater  do  not  come  into  direct 
contact.  The  steam  is  led  through  tubes  around  which  the  feed- 
water  is  forced  to  flow.  If  the  water  is  very  hard,  the  tubes  are 
liable  to  collect  scale,  which  hinders  the  operation  of  the  heater. 

73.  Economizers.  —  In  the  ordinary  steam  plant,  the  flue  gases 
pass  up  the  stack  at  a  temperature  of  about  500°  F.     This  tem- 
perature usually  will  be  higher  than  that  of  the  steam  and  water 
in  the  boiler,  since  the  latter  get  their  heat  from  the  gases.    More- 
over, the  higher  the  steam  pressure   and   its  temperature,  the 
hotter  will  be  the  flue  gas.     The  tendency  during  the  past  few 
years  has  been  to  use  higher  pressures,  which  means  a  greater  loss 
of  heat  up  the  stack  than  with  low  pressures. 


FIG.  35 


72  ENGINES   AND   BOILERS 

In  order  to  utilize  a  part  of  this  heat  that  otherwise  would  be 
wasted,  economizers  are  sometimes  installed  between  the  boiler 
and  the  stack.  The  economizer  is  simply  an  added  heating  surface 
in  the  form  of  water  tubes  about  which  the  products  of  combus- 
tion pass  on  their  way  to  the  stack.  At  best  the  feedwater  will 
be  at  a  temperature  of  only  212°  as  it  enters  the  boiler,  if  it  has 
been  heated  with  exhaust  steam  at  atmospheric  pressure.  Con- 
siderable heat  can  be  added  before  it  reaches  the  boiling  point 
when  under  high  pressure.  This  heat  is  added  in  the  econo- 
mizer. The  boiler  feedwater  is  first  pumped  to  the  economizer, 
where  it  is  heated  to  near  the  boiling  point  corresponding  to 
boiler  pressure,  and  it  then  passes  on  to  the  boiler. 

In  a  common  type  of  economizer,  the  heating  surface  is  com- 
posed of  vertical  tubes  through  which  the  water  flows  and  around 
which  the  hot  gases  pass.  These  tubes  are  kept  clean  from  soot 
by  scrapers  that  are  continually  moved  up  and  down  the 
tubes,  by  means  of  a  small  engine  or  electric  motor.  As  the 
economizer  depends  for  its  action  upon  the  extraction  of  heat 
from  the  burnt  gases,  it  follows  that  the  gases  will  be  much  cooled, 
and  if  natural  draft  be  employed,  they  may  be  cooled  enough  to  re- 
duce the  draft  to  such  an  extent  that  the  efficiency  of  the  whole 
plant  may  be  lowered.  If  forced  draft  is  used  this  objection  does 
not  hold  to  so  great  an  extent.  In  any  case,  the  economizer  offers 
some  resistance  to  the  gases,  with  a  consequent  lowering  of  the 
draft.  Whether  or  not  an  economizer  will  effect  enough  of  a 
saving  to  pay  for  itself  must  be  determined  in  each  individual 
case.  Economizers  are  often  sold  under  a  guarantee  to  add  a 
certain  percentage  to  the  efficiency  of  the  whole  plant. 

74.  Condensers.  —  After  steam  has  passed  through  an  engine  or 
turbine,  it  is  often  led  to  a  condenser,  in  which  a  pressure  con- 
siderably below  that  of  the  atmosphere  is  maintained.  The  pro- 
cess has  several  advantages  which  will  be  studied  in  more  detail 
later.  In  general,  the  decreased  back-pressure  adds  to  the  effi- 
ciency and  to  the  capacity  of  the  engine  or  turbine  to  an  extent 
more  than  sufficient  to  pay  for  the  additional  cost  of  the  condenser, 
provided  plenty  of  cool  water  is  available  for  cooling  the  con- 
denser so  as  to  condense  the  steam.  Moreover,  with  a  surface 
condenser,  the  condensed  steam  is  led  back  to  the  boiler  and  is 
thus  kept  free  from  scale-forming  materials.  This  last  factor  is  of 


BOILER   ACCESSORIES    AND    AUXILIARIES 


73 


great  importance  where  the  available  feedwater  is  poor.  Boilers 
can  be  operated  far  beyond  their  rated  capacity  if  they  are  kept 
free  from  scale.  Hence  the  saving  in  boilers  and  in  their  upkeep 
may  also  go  a  long  way  toward  paying  for  the  condenser. 

There  are  two  types  of  condensers,  one  in  which  the  cooling  or 
circulating  water  is  kept  separate  from  the  steam,  as  in  the  closed 
feedwater  heater,  and  one  in  which  the  water  (called  injection 
water)  is  mixed  with  the  steam,  as  in  the  open  feedwater  heater. 
The  former  is  called  a  surface  condenser,  and  the  latter  a  jet 
condenser. 

75.  The  Surface  Condenser.  —  Figure  36  shows  the  construc- 
tion of  a  surface  condenser.  The  exhaust  steam  enters  at  the 


5urface    Condenser? 

FIG.  36 

top  of  the  shell,  passes  around  the  tubes,  and,  after  being 
condensed,  is  pumped  out  from  the  bottom  of  the  shell.  The 
tubes  are  usually  made  of  thin  brass  or  of  special  metal,  and  extend 
between  two  tube  sheets.  At  the  outer  side  of  these  tube 
sheets  and  within  the  tubes  is  the  water  space.  The  circulating 
water  is  pumped  in  at  the  opening  shown  in  Fig.  36,  flows  through 
the  lower  half  of  the  tubes  to  the  other  end  of  the  condenser, 
and  then  flows  back  through  the  top  tubes  and  out.  This  arrange- 
ment is  called  a  two-pass  condenser.  In  some  smaller  types,  the 
water  enters  at  one  end  and  flows  out  at  the  other;  these  are 
called  one-pass  condensers. 

Occasionally  three  passes  are  made,  but  this  type  is  not  gen- 
eral. In  the  design  of  a  condenser,  care  must  be  taken  that  the 
steam,  upon  entering,  is  directed  over  the  entire  surface  of  the 
tubes  and  that  no  air  pockets  may  be  formed.  The  condensed 
steam  should  leave  that  part  of  the  condenser  where  the  circu- 
lating water  is  coldest.  Packed  joints  are  used  between  the 
tubes  and  the  sheet. 


74 


ENGINES    AND    BOILERS 


Since  water  will  absorb  and  dissolve  air  when  they  come  into 
contact,  some  air  will  be  taken  into  the  boilers  with  the 
feedwater  and  will  pass  over  with  the  steam  to  the  engine.  Air 
also  may  leak  into  the  steam  through  the  stuffing  boxes  of  the 
engine  or  turbine  when  run  condensing.  This  air  would  soon 
clog  the  condenser  and  prevent  condensation  of  the  steam  if  it 
were  not  removed.  The  air  is  pumped  out  either  with  the  con- 
densed steam  by  means  of  a  wet-air  pump  or  else  separately  by 
means  of  a  dry-air  pump. 

The  circulating  water  is  pumped  through  the  condenser  by  the 
circulating  pump.  To  maintain  a  high  vacuum,  the  circulating 
water  must  be  at  a  low  temperature  when  it  leaves  the  condenser. 
This  means  that  each  pound  of  circulating  water  can  absorb  only 
a  few  B.t.u.  Each  pound  of  the  steam  that  condenses  gives  up 
to  the  water  something  like  a  thousand  B.t.u.  It  therefore  is 
evident  that  a  large  volume  of  circulating  water  must  be  used. 
As  the  pressure  to  be  pumped  against  is  small,  a  centrifugal  pump 
is  commonly  employed  for  forcing  the  circulating  water  through 
the  condenser.  Air  pumps  are  made  both  of  the  reciprocating 
type  and  of  the  rotary  type.  The  former  are  more  common.  The 
design  of  the  air  pump  is  a  rather  difficult  problem,  since  the  air  is 
very  rare  at  a  high  vacuum,  so  that  if  the  pump 
has  much  clearance  it  will  fail  to  maintain  this 
vacuum. 

76.  The  Jet  Condenser.  —  As  stated  before, 
the  steam  and  water  mix  in  the  jet  condenser. 
This  mixture   of    condensed    steam,    injection 
water,  and  the  air  contained  in  the  steam  and 
water,  may  be  pumped  out  by  a  wet-air  pump. 
However,  the  water  is  often  allowed  to  run 
out  of  the  condenser  by  gravity  through  a  ver- 
tical pipe  thirty  feet  or  more  in  length  that  has 
its  lower  end  submerged.    The  air  is  pumped  out 
by  a  dry-air  pump.     This  arrangement  is  com- 
monly called  a  siphon  or  barometric  condenser. 
Figure  37  shows  in  section  this  latter  type  of  jet  condenser. 
The  injection  water  enters  at  A  and  runs  over  the  edges  of  trays, 
thus  exposing  a  large  surface  to  the  steam  which  enters  at  B. 
The  air  pump  sucks  the  air  out  at  the  top  through  the  pipe  C. 


FIG.  37 


BOILER   ACCESSORIES   AND    AUXILIARIES 


75 


In  this  type,  the  flow  of  the  steam,  until  it  is  condensed,  is  with 
the  air  and  opposite  to  that  of  the  water.  Such  a  condenser 
therefore  is  called  a  counter-current  condenser. 

Another  type  of  jet  condenser  in  which  the  air  pump  is  dis- 
pensed with  is  shown  in  Fig.  38.  The  air  is  carried  along  with  the 
condensed  steam  and  the  injection  water.  This  is  due  to  the  high 
velocity  of  the  water  as  it  passes  out  of  the  con- 
stricted opening  A.  This  is  called  the  injector 
or  ejector  type.  It  usually  is  furnished  with  a 
barometric  tube,  as  in  the  preceding  type. 

The  jet  condenser  is  more  compact  and  less 
expensive  than  the  surface  condenser.  If  it  is 
well  made  and  equipped  with  a  good  air  pump, 
it  will  give  a  very  high  vacuum,  but  it  mixes 
fresh  water  with  the  condensed  steam,  and  this 
may  cause  scale  if  the  mixture  is  used  for  the 
boiler  feed. 

77.  Cooling  of  Circulating  Water.  —  Since  a 
large  amount  of  cold  water  is  costly  in  some  lo- 
cations, it  is  sometimes  necessary  to  cool  the  circulating  water  so 
that  it  may  be  used  over  and  over.  This  may  be  done  by  run- 
ning it  into  a  pond  where  the  natural  evaporation  from  the  sur- 
face will  cool  it.  If  the  space  for  a  large  pond  is  not  available, 
the  evaporation  may  be  increased  by  the  use  of  spray  nozzles 
which  break  the  water  up  into  a  fine  spray,  thereby  exposing  a 
large  surface  for  evaporation.  Another  method  for  the  rapid 
cooling  of  the  circulating  or  injection  water  is  the  use  of  cooling 
towers.  The  water  is  allowed  to  trickle  down  over  a  lattice  or 
other  surface,  thus  exposing  a  large  surface  for  evaporation.  Air 
is  either  passed  up  through  the  tower  by  natural  draft,  or  blown 
through  by  means  of  fans.  Where  large  bodies  of  cold  water, 
such  as  rivers  or  lakes,  are  available,  the  circulating  water  is 
drawn  from  them  and  is  thrown  away  after  being  used.  If  the 
water  is  taken  from  lakes  or  rivers,  it  is  often  necessary  to  pass 
it  through  a  screen  to  remove  such  material  as  weeds,  drift,  and 
fish.  These  screens  require  cleaning  periodically  and  are  some- 
times made  in  the  form  of  an  endless  chain,  so  that  while  some 
sections  are  being  cleansed  others  may  be  in  use. 


CHAPTER   VI 
THE   STEAM   ENGINE 

78.  History.  —  In  1698  THOMAS  S  A  VERY  produced  the  first 
steam  engine  that  proved  to  be  commercially  successful.  It  was 
used  for  pumping  water.  The  engine  consisted  of  two  egg-shaped 
vessels,  each  of  which  connected  with  the  supply  of  water  to  be 
pumped  and  to  a  steam  boiler.  In  its  operation  steam  was  ad- 
mitted to  one  of  the  vessels,  and  when  it  was  full,  connection 
with  the  boiler  was  cut  off.  Cold  water  was  then  sprayed  on  the 
outer  surface  of  the  vessel,  causing  the  steam  inside  to  condense 
and  to  form  a  partial  vacuum.  This  vacuum  opened  a  valve  in 
the  pipe  leading  to  the  well  and  sucked  water  into  the  vessel. 
Steam  was  then  turned  on  again,  and  the  pressure  forced  the 
water  out  of  the  vessel  through  the  delivery  pipe.  When  the 
water  had  all  been  forced  out,  the  process  described  above  was 
repeated.  The  engine  was  operated  so  that  while  one  cylinder 
was  forcing,  the  other  was  sucking  water.  This  engine  is  the 
same  in  principle  as  the  modern  pulsometer. 

Since  the  steam  in  the  Savery  engine  came  in  direct  contact 
with  the  water  during  the  forcing  stroke,  there  was  much  loss 
of  steam  by  useless  condensation.  DENIS  PAPIN,  in  1705,  made 
an  improvement  on  the  Savery  engine  by  making  the  steam  vessel 
of  cylindrical  shape  and  separating  the  steam  and  water  by  a 
floating  piston,  thereby  preventing  a  part  of  the  unnecessary  con- 
densation. 

About  1711,  there  came  into  use  a  machine  that  was  known 
as  the  Newcomen  engine.  THOMAS  NEWCOMEN,  with  the  aid  of 
JOHN  GALLEY,  and  with  certain  ideas  from  Papin,  made  his  engine 
with  a  vertical  cylinder  into  which  a  piston  was  fitted  from  the  upper 
end.  The  cylinder  was  placed  directly  above  the  boiler  and  con- 
nected with  it.  Steam  was  admitted  to  the  cylinder  by  the  open- 
ing of  a  valve  placed  between  the  boiler  and  the  cylinder.  The 
piston  was  connected  to  a  pump  through  a  walking  beam,  one 
end  of  the  walking  beam  connecting  to  the  pump  rod  and  the 
other  to  the  piston  rod.  The  beam  was  so  counterbalanced  that 
it  took  but  little  steam  pressure  to  force  the  piston  up.  Steam 
was  generated  at  about  atmospheric  pressure,  and  as  the  piston 

76 


THE    STEAM    ENGINE  77 

moved  up  the  steam  valve  was  opened  and  steam  filled  the  space 
beneath  it.  When  at  the  top  of  its  stroke,  water  was  sprayed 
into  the  cylinder,  condensing  the  steam  and  forming  a  partial 
vacuum.  This  vacuum  under  the  piston  allowed  the  atmos- 
pheric pressure  from  above  to  force  the  piston  down. 

While  the  Newcomen  engine  was  an  improvement  over  the 
Savery  engine,  it  was  very  wasteful  of  steam  because  the  cylinder 
was  cooled  by  the  spray  of  water  on  each  downward  stroke.  Much 
condensation  of  steam  occurred  in  heating  up  the  cylinder  walls 
on  each  upward  stroke.  While  repairing  one  of  these  Newcomen 
engines,  JAMES  WATT  conceived  ways  in  which  it  might  be  im- 
proved. Patents  covering  these  ideas  were  granted  him  in  1769. 
Watt's  chief  aim  was  to  keep  the  cylinder  walls  as  hot  as  the 
incoming  steam  at  all  times  and  thereby  prevent  the  initial  con- 
densation that  rendered  the  older  engine  so  inefficient.  This  high 
cylinder  temperature  was  to  be  maintained,  first,  by  condensing 
the  steam  in  a  vessel  away  from  the  cylinder,  and  second,  by  a 
steam  jacket  placed  around  the  cylinder  walls.  Lagging  was  also 
to  be  placed  around  the  outside  of  the  cylinder  to  keep  down  the 
heat  lost  to  the  outside  air.  In  the  previous  engines,  the  piston 
was  kept  tight  by  a  kind  of  water  seal  on  top  of  the  piston.  Watt 
used  fibrous  packing  and  tallow  to  keep  the  piston  tight  and 
saved  heat  that  previously  was  lost  to  the  water  above  the  piston. 
In  the  operation  of  his  engine  Watt  found  it  necessary  to  remove 
the  air  from  the  condenser  and  so  he  equipped  the  condensers 
with  air  pumps. 

While  it  is  seen  that  Watt  is  not  the  inventor  of  the  steam 
engine,  yet  it  must  be  admitted  that  he  did  more  to  advance 
its  development  than  any  other  one  man.  Up  to  the  time  of 
Watt,  the  steam  engine  was  used  almost  exclusively  for  pumping 
water  in  collieries,  but  he  applied  it  to  the  driving  of  other  forms 
of  machinery.  After  many  hardships  and  discouragements,  Watt 
at  last  was  able  to  produce  his  engine  in  large  numbers.  The  engine 
became  increasingly  popular,  and  we  may  say  that  the  era  of  our 
present  industrial  development  started  at  the  time  of  James  Watt. 
In  applications  for  his  patents  Watt  advocated  the  use  of  high- 
pressure  steam  from  which  work  could  be  obtained  by  using  it 
expansively,  but  in  the  actual  construction  of  his  engines  he  never 
used  pressures  much  above  that  of  the  atmosphere. 

Since  the  time  of  Watt,  various  improvements  have  been  made 


78  ENGINES   AND   BOILERS 

in  the  steam  engine.  The  mechanical  construction  has  been  bet- 
tered, the  valve  mechanism  improved,  compounding  adopted,  and 
the  steam  pressures  greatly  increased.  While  its  thermal  effi- 
ciency may  not  be  as  high  as  some  forms  of  internal-combustion 
engines,  the  steam  engine  is  very  reliable. 

79.  The  Plain  Slide-valve  Engine.  —  Figure  39  shows  in  ver- 
tical and  horizontal  sections  the  parts  of  a  simple  steam  engine. 
Its  action  is  as  follows :  Steam  comes  from  the  boiler  through  the 
steam-pipe,  and  after  passing  the  throttle  valve,  enters  the  steam 
chest.  The  valve,  driven  by  an  eccentric  on  the  shaft,  moves 
backward  and  forward  on  the  valve  seat,  uncovering  alternately 
the  two  steam  ports.  When  a  steam  port  is  uncovered  by  the 
valve,  the  steam  flows  through  the  port  into  the  cylinder  and  by 
its  pressure  moves  the  piston  in  the  cylinder.  In  Fig.  39,  the 
left  steam  port  is  shown  partly  open  and  the  steam  is  then  push- 
ing the  piston  to  the  right.  At  the  same  time  that  the  steam  is 
forcing  the  piston  to  the  right,  the  valve  has  uncovered  the  right 
port,  so  that  the  steam  on  the  right  of  the  piston  may  escape  to 
the  exhaust  pipe. 

This  motion  of  the  piston  is  transmitted  by  the  piston  rod  to 
the  cross-head  and  from  this  through  the  wrist  pin  and  the  con- 
necting rod  to  the  crank.  The  reciprocating  motion  of  the  pis- 
ton is  transformed  by  the  connecting  rod  and  crank  into  rotary 
motion  of  the  shaft.  The  power  generated  in  the  cylinder  usually 
is  taken  from  the  shaft  by  a  belt  on  the  flywheel  or  by  an  electric 
generator  coupled  to  the  shaft.  The  valve  is  made  to  move  so 
that  when"  steam  is  being  admitted  to  one  end  of  the  cylinder  it 
is  being  exhausted  from  the  other.  As  the  incoming  steam  is  at 
a  much  higher  pressure  than  the  exhaust,  there  is  a  resultant 
force  pushing  the  piston  in  the  direction  of  the  outgoing  steam. 

That  end  of  the  cylinder  farthest  from  the  crank  is  called  the 
head-end^  and  the  end  nearest  the  crank  the  crank-end.  With 
the  piston  at  the  extreme  left  of  its  travel,  the  crank  will  be  in 
a  direct  line  between  the  cylinder  and  the  shaft.  While  the  force 
on  the  crank  pin  may  be  large,  there  is  no  turning  effort.  The 
crank  is  then  said  to  be  on  head-end  dead  center.  With  the  pis- 
ton at  the  extreme  right  of  its  travel  the  crank  is  on  crank- 
end  dead  center.  When  the  engine  is  running,  if  the  crank  rises, 
as  the  piston  leaves  the  head-end  dead  center  (i.e.,  if  in  Fig.  39 


THE    STEAM   ENGINE 


79 


the  crank  moves  in  a  clockwise  direction),  the  engine  is  said  to 
be  running  over.  If  the  crank  moves  in  the  opposite  direction, 
the  engine  is  said  to  be  running  under. 


FIG.  39 


The  stroke  of  the  engine  is  the  distance  the  piston  travels  in 
half  a  revolution.  It  is  equal  to  twice  the  length  of  the  crank. 
On  the  head-end  stroke,  the  piston  moves  from  head-end  to 
crank-end  dead  center,  and  the  reverse  motion  takes  place  on 
the  crank-end  stroke. 


80  ENGINES   AND    BOILERS 

80.  Parts  of  the  Steam  Engine.  —  CYLINDER.  Steam-engine 
cylinders  are  made  of  cast  iron  and  the  bore  is  carefully  machined. 
With  proper  lubrication,  the  surface  exposed  to  wear  acquires  a 
high  polish  and  the  metal  is  worn  away  very  slowly.  The  ports 
are  cored  in  the  casting  and  are  not  finished  except  along  the 
edges  at  the  valve  seat.  At  each  end  of  the  cylinder  the  diameter 
is  made  slightly  larger.  This  enlarged  part  is  called  the  counter- 
bore.  It  should  extend  far  enough  so  that  the  piston  ring  comes 
to  its  edge,  in  order  that  a  shoulder  may  not  be  worn  in  the  cylin- 
der wall.  The  counterbore  also  serves  a  purpose  when  the  cylinder 
has  to  be  rebored,  since  the  boring  machine  may  be  set  on  the 
counterbore,  which  will  not  be  worn  away,  and  thus  the  alignment 
need  not  be  lost.  On  some  larger  engines  used  in  marine  service, 
a  thin  inner  shell  or  liner  is  placed  in  the  cylinder,  so  that  it  may 
be  replaced  without  changing  the  cylinder  when  it  becomes  worn. 

The  cylinder  head  is  bolted  to  the  cylinder,  and  the  joint  is 
made  steam-tight  by  means  of  a  gasket  or  a  ground  joint.  In  the 
smallest  engines  the  cylinder  is  cast  as  a  part  of  the  frame,  but 
ordinarily  it  is  cast  separately  and  bolted  to  the  frame. 

In  large  engines  where  high  efficiency  is  desired  the  cylinder 
may  have  a  steam  jacket.  The  heads  may  or  may  not  be  jack- 
eted. In  any  case  the  cylinder  is  covered  by  some  non-conductor 
of  heat,  called  lagging.  With  proper  lagging  very  little  heat  is 
lost  by  radiation. 

Unless  the  exhaust  valves  are  so  placed  that  any  water  in  the 
cylinder  can  drain  to  them,  it  is  necessary  to  tap  the  bottom  of 
the  counterbore  at  each  end  of  horizontal  cylinders  and  place  a 
drain  cock  there.  These  cocks  are  opened  when  the  engine  is 
warming  up  so  that  the  condensed  steam  will  not  be  caught  when 
the  piston  comes  to  the  end  of  the  stroke.  As  water  is  incom- 
pressible, its  presence  in  too  large  a  quantity  would  cause  the 
breaking  or  straining  of  some  part.  Sometimes  automatic  relief 
valves  are  placed  at  the  ends  of  the  cylinders  to  take  care  of  any 
water  that  may  get  into  the  cylinders.  Each  end  of  the  cylinder 
is  tapped  for  the  connection  of  an  indicator. 

VALVES.  The  subject  of  valves  will  be  taken  up  in  detail  in 
Chapter  VIII. 

PISTON.  Pistons  for  stationary  engines  are  made  of  cast  iron. 
The  piston  is  turned  to  a  slightly  smaller  diameter  than  the 


THE    STEAM   ENGINE  81 

cylinder,  and  leakage  of  steam  past  it  is  prevented  by  the  use  of 
piston  rings  which  fit  into  grooves  cut  around  the  piston.  In  the 
small  sizes  the  rings  are  made  in  one  piece  slightly  larger  than 
the  diameter  of  the  cylinder.  A  piece  is  then  cut  out  of  each 
ring,  and  they  are  snapped  into  the  groove  in  the  piston. 
Because  of  their  elasticity,  they  spring  out  and  make  contact  with 
the  cylinder  walls.  When  worn  they  may  be  replaced  by  new 
rings.  On  larger  pistons  the  rings  are  built  up  in  sections  and 
are  pushed  out  against  the  cylinder  wall  by  springs  placed  under 
them. 

On  small  engines  the  pistons  are  cast  in  one  piece,  and  usually 
are  hollow,  to  make  them  as  light  as  possible.  In  the  larger 
sizes,  they  are  built  up  of  two  or  more  pieces.  In  vertical  engines 
and  locomotives,  the  pistons  sometimes  are  made  dished  to  a 
slight  extent.  The  dishing  may  be  to  add  strength,  to  shorten 
the  length  of  the  engine  a  small  amount,  or  to  facilitate  drainage. 

PISTON  ROD.  The  piston  rod  connects  the  piston  to  the  cross- 
head.  It  is  made  of  steel,  and  the  connection  must  be  such  that 
it  will  be  always  tight.  If  a  little  play  is  allowed,  a  bad  knock 
develops  that  rapidly  grows  worse.  In  horizontal  engines,  a  tail- 
rod  sometimes  extends  from  the  piston  out  through  the  head-end 
cylinder  head,  and  its  outer  end  is  carried  by  a  slipper  on  a  guide. 
This  arrangement  allows  the  weight  of  the  piston  to  be  carried 
by  the  slipper  and  the  cross-head  and  lessens  the  wear  on  the 
piston  and  the  cylinder. 

STUFFING  Box.  The  joint  between  the  piston  rod  and  crank- 
end  cylinder  head  is  made  tight  by  a  stuffing  box.  The  packing 
used  on  low-pressure  engines  in  this  box  may  be  fibrous,  but  with 
high  steam  pressure  a  metallic  packing  is  commonly  used.  A  good 
packing  should  keep  the  joint  steam-tight  and  at  the  same  time 
give  but  little  friction  on  the  rod.  Wear  in  the  cylinder,  improper 
adjustment  of  the  cross-head,  poor  alignment,  or  a  pitted  or  scored 
rod,  may  cause  excessive  wear  on  the  packing.  It  is  then  diffi- 
cult to  keep  it  steam-tight. 

CROSS-HEAD.  The  cross-head  with  the  cross-head  pin  or  wrist 
pin  forms  the  connection  between  the  piston  rod  and  the  con- 
necting rod.  The  cross-head  is  made  to  move  in  a  straight  line 
by  guides  on  the  frame  of  the  engine.  The  two-guide  type  shown 
in  Fig.  39  is  the  most  common,  although  four-guide  and  one- 


82  ENGINES   AND   BOILERS 

guide  types  are  used  occasionally.  The  slipper  type  also  is  often 
used;  in  this  the  cross-head  takes  the  form  of  a  slipper  which 
slides  on  the  flat  surface  of  one  guide.  In  all  types  the  wear 
between  the  cross-head  and  guides  is  taken  up  by  the  adjust- 
ment of  the  wedge-shaped  slippers  or  by  the  use  of  shims.  In  a 
few  steam  engines  and  most  gas  engines  a  trunk  piston  is  used. 
In  this  type  the  piston  itself  acts  as  cross-head  and  carries  the 
wrist  pin.  With  this  latter  arrangement,  the  engine  is  single 
acting,  i.e.  the  steam  acts  only  on  the  head  of  the  cylinder  and  on 
the  face  of  the  piston. 

CONNECTING  ROD.  The  connecting  rod  connects  the  wrist  pin 
to  the  crank  pin.  It  is  alternately  in  compression  and  tension, 
and  usually  is  made  of  steel.  On  high-speed  engines,  considerable 
bending  stress  may  be  developed  in  the  connecting  rod  on  account 
of  its  fling,  and  hence  its  cross-section  is  usually  rectangular  or 
I-section.  On  slow-speed  engines  this  bending  stress  is  small, 
and  the  rod  is  made  circular  in  section.  Brass  or  other  metal 
bearing-pieces  called  brasses  are  used  at  the  bearing  points  on  the 
pins.  These  brasses  are  often  babbited.  As  wear  occurs  ad- 
justment must  be  made  to  take  it  up.  If  the  rod  is  shortened 
in  taking  up  the  wear,  the  end  on  which  such  adjustment  is 
made  is  said  to  have  an  open  stub-end.  If  the  rod  is  length- 
ened by  an  adjustment  at  one  end,  that  end  is  called  a  closed 
stub-end.  In  Fig.  39,  the  end  at  the  crank  pin  is  of  the  marine 
stub  type,  which  is  used  on  all  center-crank  engines.  In  this 
last  type  the  wear  is  taken  up  by  removing  liners,  and  in  the 
former  types  usually  by  means  of  wedges  v/hich  move  the  brasses. 

CRANK.  The  crank  pin  is  made  of  steel,  and  it  may  be  a  part 
of  the  same  forging  as  the  crank  and  shaft,  or  it  may  be  set  into 
crank  discs  which  are  keyed  to  the  shaft.  When  the  pin  is  placed 
between  two  crank  discs,  as  in  Fig.  39,  we  have  a  center-crank 
engine;  when  it  overhangs  one  crank  disc,  we  have  a  side-crank 
engine. 

COUNTERBALANCE.  The  crank  pin  and  half  of  the  connecting 
rod  usually  are  considered  as  rotating  parts  and  must  be  coun- 
terbalanced to  make  the  engine  run  smoothly.  Furthermore  the 
piston,  piston  rod,  cross-head,  and  half  of  the  connecting  rod  have 
a  reciprocating  motion.  It  takes  a  large  force  to  start  and  stop 
them  on  each  stroke.  Unless  they  are  counterbalanced  the  whole 


THE    STEAM   ENGINE  83 

engine  will  vibrate  on  the  foundation.  While  it  is  impossible  to 
counterbalance  exactly  both  the  rotating  and  the  reciprocating 
parts  at  the  same  time,  yet  it  can  be  partially  done.  The  proper 
sized  counterbalance  or  counterweight  (Fig.  39),  sometimes  made 
of  lead  but  usually  of  iron,  is  put  on  to  give  smooth  running. 

SHAFT.  Engine  shafts  usually  are  made  of  steel.  As  explained 
above,  they  are  either  forged  to  make  the  crank  and  crank  pin 
integral  parts  of  the  shaft,  or  else  the  crank  is  keyed  to  the  shaft. 
In  addition  to  the  key,  a  shrunk  fit  sometimes  is  used,  or  the 
crank  disc  may  be  pressed  on  by  hydraulic  pressure. 

BEARINGS.  With  a  side-crank  engine  there  is  one  main  bear- 
ing, and  with  a  center-crank  engine,  there  are  two.  The  weight 
of  the  flywheel  may  be  carried  partly  by  an  outer  bearing  called 
an  outboard  bearing.  There  will  be  wear  on  the  main  bearing 
in  a  vertical  direction  on  account  of  the  weight  of  the  flywheel 
and  of  the  rotating  parts,  and  there  will  be  wear  in  a  horizontal 
direction  on  account  of  the  thrust  from  the  piston.  Many  main 
bearings  are  made  up  of  four  parts,  the  cap,  the  bottom  part 
which  takes  the  vertical  wear,  and  two  side-pieces  which  take 
the  horizontal  wear.  The  latter  are  called  quarter-boxes.  These 
parts  may  be  adjusted  separately. 

FLYWHEEL.  The  turning  moment  on  the  crank  varies  at  differ- 
ent parts  of  the  stroke.  At  dead  center  it  is  zero.  In  order  to 
keep  the  shaft  turning  at  approximately  the  same  speed  at  all 
times  in  the  revolution,  a  flywheel  is  put  on  the  shaft.  This  acts 
to  store  up  and  give  out  energy  at  the  proper  times,  thereby  keep- 
ing the  angular  velocity  approximately  uniform.  The  flywheel  may 
carry  the  belt  or  there  may  be  a  separate  belt  wheel  in  addition 
to  the  flywheel,  in  which  case  the  latter  is  sometimes  called  a 
balance  wheel.  The  flywheel  commonly  is  made  of  cast  iron. 
In  the  smaller  sizes  it  is  cast  in  one  piece,  but  in  the  larger  sizes, 
it  is  cast  in  sections.  Great  care  must  be  taken  in  its  manufac- 
ture, since  a  crack  may  cause  a  disastrous  accident. 

ECCENTRIC.  An  eccentric  is  placed  on  the  shaft  or  governor  arm 
to  drive  the  valve.  It  is  encircled  by  the  eccentric  strap,  which 
is  connected  to  the  valve  by  means  of  the  eccentric  rod  and 
valve  stem.  This  is  really  a  substitute  for  a  crank  and  connect- 
ing rod  and  gives  to  the  valve  a  motion  similar  to  that  of  the 
piston. 


84  ENGINES   AND   BOILERS 

FRAME.  The  frame  is  made  of  cast  iron  in  stationary  engines, 
and  the  better  engines  have  the  heavier  frames.  The  greater  the 
weight  of  frame  the  more  smoothly  the  engine  will  run,  other 
things  being  equal.  On  some  small  high-speed  engines  there  is  a 
cast-iron  sub-base  placed  between  the  frame  and  the  foundation. 
Frames  are  given  different  names  according  to  their  shape  and 
the  cylinder  arrangement. 

FOUNDATION.  Foundations  usually  are  made  of  brick  or  con- 
crete. The  latter  is  now  the  more  common.  The  frame  is  fas- 
tened to  the  foundation  by  anchor  bolts.  The  foundation  should 
be  quite  massive  and  should  rest  on  soil  that  is  firm  enough  to 
carry  the  weight  of  the  engine  and  foundation  without  settling. 

81.  Piston  Displacement.  —  The  volume  the  piston  displaces 
in  moving  from  one  dead  center  to  the  other  is  called  the  piston 
displacement.    It  is  commonly  expressed  in  cubic  feet.    The  head- 
end piston  displacement  is  equal  to  the  length  of  stroke  in  feet 
times  the  area  of  the  piston  in  square  feet.    The  crank-end  piston 
displacement  is  the  area  of  the  piston  minus  the  area  of  the  cross- 
section  of  the  piston  rod,  times  the  stroke.  . 

The  size  of  an  engine  is  given  in  inches,  the  diameter  of  the 
cylinder  bore  first  and  the  length  of  the  stroke  second,  e.g.  an 
18//X24'/  engine  is  one  whose  cylinder  is  18"  in  internal  diameter 
and  whose  stroke  is  24".  The  size  of  a  compound  engine  is  given 
by  the  diameter  of  the  high-pressure  cylinder,  the  diameter  of  the 
low-pressure  cylinder,  and  the  stroke,  as  10"X18'/X24'/.  If  the 
volume  is  calculated  in  cubic  inches,  remember  to  change  tc  cubic 
feet  in  giving  the  piston  displacement. 

82.  Clearance.  —  When  the  piston  is  at  the  extreme  end  of  its 
travel,  there  will  be  some  volume  back  of  it,  because  it  is  neces- 
sary to  have  a  little  space  in  which  to  take  up  the  wear  on  the 
connecting  rod  brasses  and  to  allow  for  unequal  expansion  of 
parts  as  the  engine  heats  up,  and  because  of  the  space  in  the 
ports.    This  volume,  the  larger  part  of  which  is  often  the  volume 
of  the  ports,  is  called  clearance.    Clearance  is  expressed  as  a  per- 
centage of  the  piston  displacement.    Thus,  when  we  say  that  the 
head-end  clearance  of  an  engine  is  4.5  percent,  we  mean  the 
volume  back  of  the  piston  when  the  engine  is  on  head-end  dead 
center  is  .045  times  its  head-end  piston  displacement. 

To  determine  the  clearance  of  an  engine,  first  place  the  engine 


THE    STEAM   ENGINE  85 

on  the  dead  center  of  the  end  for  which  the  clearance  is  to  be 
measured.  Second,  disconnect  the  valve  from  the  eccentric 
rod,  move  it  so  that  the  port  for  that  end  is  closed,  and  block  it 
up  in  that  position.  It  is  necessary  to  disconnect  the  valve  be- 
cause the  port  is  usually  open  a  small  amount  when  the  engine 
is  on  dead  center.  Third,  pour  water  into  the  opening  for 
attaching  the  indicator  cock  until  the  clearance  space  is  full  of 
water.  If  the  valve  is  placed  at  the  top  of  the  cylinder,  as  is 
the  case  with  horizontal  Corliss  engines,  remove  the  valve  and 
pour  the  water  into  the  port.  Having  recorded  the  amount  of 
water  poured  in,  and  having  made  correction  for  leakage,  the 
volume  of  wrater  necessary  to  fill  the  clearance  space  is  computed. 
This  volume  divided  by  the  piston  displacement  for  that  end  gives 
the  clearance.  A  rough  method  of  computing  clearance  from  the 
indicator  diagram  is  given  in  §  89. 

83.  Steam  Back  of  Piston  during  Stroke.  —  The  weight  of  the 
dry  steam  back  of  the  piston  may  be  computed  for  any  percent 
of  the  stroke  if  we  know  the  clearance,  the  size  of  the  engine, 
and  the  steam  pressure  at  that  percent  of  the  stroke.     Add  the 
percent  of  the  stroke  to  the  percent  of  clearance.     Multiply  this 
by  the  piston  displacement,  and  divide  by  100.     The  result  is 
the  volume  of  steam  back  of  the  piston  at  the  given  percent  of 
stroke.    By  means  of  the  indicator,  the  pressure  may  be  determined 
for  the  same  position.     The  density  of  steam  may  be  read  from 
the  steam  tables  for  that  particular  pressure.     The  product  of 
this  density  and  the  volume  back  of  the  piston  gives  the  weight 
of  the  steam  there. 

EXAMPLE.  What  is  the  weight  of  dry  steam  back  of  the  piston  at  27% 
of  the  crank-end  stroke  of  a  14"X16"  engine?  The  engine  has  a  2"  rod  and 
the  crank-end  clearance  is  6.3%.  The  crank-end  indicator  card  is  at  hand. 

SOLUTION.  Measure  the  pressure  from  the  card  at  27%  of  the  stroke. 
Suppose  this  pressure  is  115  pounds  gage,  and  the  atmospheric  pressure  is 
14.5  pounds  per  square  inch.  The  absolute  pressure  is  then  115  +  14.5  =  129.5 
pounds  per  square  inch.  The  density  of  dry  saturated  steam  at  this  pres- 
sure is,  from  the  steam  tables,  .2887  The  piston  displacement  is  seen  to  be 
16(7r72  — TT)/  1728  =  1.398  and  the  volume  back  of  the  piston  is  therefore  equal 
to  (.27 +  .063)  XI. 398  =  .466  cubic  feet.  The  weight  of  dry  steam  is  then 
.2888  X  .466  = .  1 342  pound. 

84.  The  Indicator  and  Its  Purposes.  —  The  steam-engine  indi- 
cator was  first  used  by  JAMES  WATT,  who  invented  it.     Since  his 
time  it  has  been  perfected  and  is  now  very  extensively  used. 


86  ENGINES    AND    BOILERS 

The  indicator  records  on  paper  a  line  showing  the  relation  between 
the  pressure  in  the  cylinder  and  the  movement  of  the  piston. 
The  diagram  or  card  produced  is  of  value  in  setting  the  valves,  in 
computing  the  horsepower  developed  in  the  cylinder,  and  in  mak- 
ing analyses  of  the  operation  of  the  engine.  A  description  of  the 
mechanism  of  the  indicator  will  not  be  given  here.  When  we 
speak  of  the  scale  of  spring  of  an  indicator,  we  do  not  mean  the 
actual  scale  of  the  spring  used  in  the  indicator,  but  rather  the 
relation  between  the  movement  of  the  pencil  on  the  indicator  dia- 
gram and  the  pressure  that  causes  it.  That  is,  a  60-pound  indicator 
spring  is  one  that  gives  a  pencil  movement  of  one  inch  for  each 
60  pounds  per  square  inch  increment  of  pressure  in  the  cylinder. 
The  heavy  lines  of  Fig.  40  show  a  representative  diagram  as 
it  comes  from  the  indicator.  It  is  seen  that  the  diagram  is  a 
closed  irregular  curve  with  a  straight  line  underneath.  The 
Cuf.off  straight  line  is  the  atmospheric- 

pressure  line,  which  is  used  for 
reference    in    measuring    pres- 
._  ____  +  ^e/eeje  sures.      The    ordinates   of  the 
points  on  the  curve   show  to 
*  some  scale  the  pressures  in  the 


jpIGt  40  cylinder,  and  the  abscissas  the 

movement  of  the  piston.    Start- 

ing at  the  upper  left-hand  corner  of  Fig.  40,  we  have  the  pressure 
back  of  the  piston  when  it  is  on  dead  center.  As  the  piston 
moves  forward,  the  pressure  changes  as  shown  by  the  upper  curved 
line  of  the  diagram.  It  is  seen  that  the  pressure  drops  but  little 
for  the  first  part  of  the  stroke,  and  later  more  rapidly,  until  at  the 
end  of  the  stroke  it  is  nearly  down  to  that  of  the  atmosphere. 
On  the  backward  stroke,  the  pressure  remains  nearly  constant 
until  near  the  end,  when  it  rapidly  rises  to  the  initial  point. 

85.  Events  of  Stroke.  —  There  are  four  events  of  the  stroke. 
1.  Admission  occurs  when  the  valve  uncovers  the  port  and  allows 
the  steam  to  enter  the  cylinder.  2.  Cut-off  takes  place  when 
the  valve  closes  the  port  and  prevents  any  more  steam  from 
entering.  3.  At  release,  the  valve  uncovers  the  port  and  allows 
the  steam  to  escape  to  the  exhaust.  4.  Compression  occurs  when 
the  valve  again  closes  the  port  and  prevents  any  more  steam 
from  leaving  the  cylinder  for  the  remainder  of  the  stroke. 


THE    STEAM   ENGINE  87 

It  is  seen  that  the  steam  enters  the  cylinder  from  admission  to 
cut-off,  and  that  the  steam  thus  let  in  expands  and  does  work 
on  the  piston  from  cut-off  to  release.  From  release  to  compression, 
the  used  steam  is  being  exhausted  from  the  cylinder.  The  steam 
that  is  caught  when  the  exhaust  port  closes  is  compressed  into  the 
clearance  space  during  the  time  from  compression  to  admission. 

86.  Location  of  Events  on  Diagram.  —  After  a  little  practice 
the  events  may  be  quite  accurately  located  on  the  ordinary  in- 
dicator diagram  or  card.     In  Fig.  40    it  is  seen  that  the  upper 
line  drops  down  and  that  there  is  a  point  of  inflection  in  this 
curve.     This  is  the  point  of  cut-off.     This  point  of  inflection  is 
easily  detected  by  drawing  a  continuation  of  the  two  curves  as 
shown  by  the  dotted  lines  at  the  point  of  cut-off.     Release  occurs 
at  the  next  point  of  inflection;    it  may  be  located  in  a  manner 
similar  to  that  in  which  we  located  the  cut-off.     At  compression 
there  is  no  point  of  inflection;  therefore  its  proper  location  is  diffi- 
cult.    The  exhaust  valve  ordinarily  closes  rather  slowly,  and  the 
passageway  for  steam  being  small  when  it  is  nearly  shut,  the  pres- 
sure may  start  to  rise  even  before  the  valve  is  completely  closed. 

A  common  error  is  that  of  taking  the  point  of  compression 
too  low  on  the  compression  curve.  Admission  occurs  where  the 
compression  curve  stops  and  the  straight  line  starts. 

To  determine  the  percentage  of  stroke  at  the  different  events, 
draw  the  two  end-ordinates.  The  distance  between  these,  /  in 
Fig.  40.  represents  the  length  of  stroke  to  some  scale.  The  dis- 
tances from  the  left  end-ordinate  to  the  events  shown  by  a,  b,  e, 
and  d,  divided  by  the  distance  /,  give  the  ratios  of  the  stroke 
at  these  events,  and  the  percentages  of  stroke  are  those  ratios 
multiplied  by  100. 

87.  Equation  of  Expansion  and  Compression  Curves.  —  There 
is  a  definite  relation  between  pressure  and  volume  during  expan- 
sion and  compression.     The  equation  pvn  =  p\v\n  =  p2V2n  expresses 
this  relation,  where  p  is  the  absolute  pressure  on  these  curves, 
v  is  the  volume  back  of  the  piston  (including  the  clearance  space), 
and  n  is  some  constant  exponent  for  each  individual  curve.     An 
analysis  of  many  cards  shows  that  n  is  sometimes  a  little  less 
than  1  and  sometimes  a  little  larger  than  1.     For  rough  calcu- 
lations, it  may  be  considered  equal  to  1.     With  this  assumption, 
the  equation  of  the  expansion  and  compression  curves  is  pv=C. 


ENGINES   AND   BOILERS 


This  is  the  equation  of  the  equilateral  hyperbola.  Certain  geo- 
metric facts  about  this  curve  are  of  importance  to  us.  In  Fig.  41, 
let  us  choose  any  two  points  on  an  equilateral  hyperbola:  E, 
whose  coordinates  are  (pi,  vi),  and  H,  whose  coordinates  are 
(P2,  ^2)-  We  shall  show  that  the  line  of  the  diagonal  CA  of  the 
rectangle  EC  HA  drawn  through  these  points,  passes  through  the 
origin.  In  the  triangles  OAB  and  OCD,  OB  =  Vi,  BA=p2, 
OD=vz,  and  DC  =pi.  Since  their  sides  are  parallel,  these  two 
triangles  are  similar.  Hence  we  have 


yo/ume    t/j   Ct/jb/c 


which  satisfies  the  equation  of  the  curve. 

It  follows  that  we  can  construct  the  entire  curve  if  one  point  E 

on  it  is  given.     We  can  locate  other  points  on  it  in  the  following 

manner.  Through  the  point  E 
draw  horizontal  and  vertical 
lines.  Choose  some  point  C  on 
the  horizontal  line,  and  draw 
the  line  OC  through  the  origin. 
Drop  a  vertical  line  from  C, 
and  draw  a  horizontal  line 

jj    x/x  "      '  '  'f%r  through  A,  the  point  where  the 

line  OC  cuts  the  vertical  line 
through  E.  The  intersection  H 
of  the  vertical  line  through  C 

and  the  horizontal  line  through  A  is  a  point  on  the  curve.     Other 

points  may  be  found  in  the  same  manner. 

Another  geometric  fact  of  value  is  that  the  length  of  FE  is 

equal  to  HG  on  a  line  drawn  through  the  points  E  and  H.     This 

may  be  proved  readily  from  Fig.  41.     The  area  under  the  curve 

from  E  to  H,  i.e.  BEHD,  may  be  determined  by  integration. 

The  increment  of  this  area  has  dimensions  p  and  dv,   and  dA 

is  equal  to  pdv.     Hence  the  total  area  is 

A  =  \  dA  =  \    pdv, 
but  since  p\v\  =  p%V2  =  pv,  we  have 


and 


THE    STEAM   ENGINE 


89 


88.  Hypothetical  Indicator  Diagram.  —  Diagrams  are  some- 
times constructed  on  the  hypotheses  that  the  expansion  and  com- 
pression curves  are  equilateral  hyperbolas,  that  the  pressure  dur- 
ing the  admission  of  steam  from  the  end  of  the  stroke  to  cut-off 
is  constant,  and  that  the  back  pressure  is  constant  up  to  the 
point  of  compression.  Such  a  diagram  will  be  called  the  hypo- 
thetical diagram.  This  diagram  is  also  sometimes  called  the 
theoretical,  or  ideal,  or  conventional  diagram. 

The  construction  of  such  a  diagram,  shown  in  Fig.  42,  is  car- 
ried out  as  follows.  First  choose  a  suitable  length  /  for  the  dia- 
gram, and  draw  in  the  atmospheric-pressure  line.  Next  choose 

a    scale    of    pressures,  ^ ^ 

and  draw  the  volume  ~~~  e  '  •  —  -  -« 

-. 

axis  at  a  distance  c,  the 
atmospheric  pressure, 
below  the  atmospheric 
line.  Draw  the  pres- 
sure axis  at  a  distance 
from  the  point  F  of  the 
diagram  equal  to  the 
ratio  of  clearance  times 
the  length /of  the  dia- 
gram. Measure  up  from 
the  atmospheric-pres- 
sure line  the  initial  steam  pressure,  and  draw  the  steam-admis- 
sion line  FA.  The  length  of  FA  is  equal  to  the  ratio  of  cut-off 
times  the  length  /.  From  the  point  A,  construct  an  equilateral 
hyperbola  as  in  §  87.  The  length  e  is  equal  to  the  ratio  of  release 
times  /.  From  B,  the  point  of  release,  draw  a  line  to  the  end  of 
the  diagram  at  C.  The  distance  h  from  C  to  the  atmospheric- 
pressure  line  represents  the  back  pressure.  From  C  to  D  draw  the 
back-pressure  line  parallel  to  the  atmospheric  line.  From  D, 
the  point  of  compression,  construct  an  equilateral  hyperbola  to 
the  point  E,  whose  distance  a  from  the  end  of  the  diagram  is  the 
admission  distance.  Connect  E  and  F  by  a  straight  line. 

The  actual  diagram  may  vary  considerably  from  the  diagram 
just  constructed.  The  speed  of  the  engine,  the  throttling  of 
steam  in  the  ports  and  by  the  valve,  the  condensation  in  the 
cylinder,  etc.,  affect  the  form  of  the  actual  diagram,  which  may 
resemble  the  dotted  diagram  FGHCIJF  in  Fig.  42.  If  the  en- 


in  £tsto/c  feef 
FIG.  42 


90 


ENGINES   AND    BOILERS 


FIG.  43 


gine  exhausts  into  a  condenser  in  which  a  vacuum  is  maintained, 
the  back-pressure  line  will  fall  below  the  atmospheric-pressure  line. 

89.  Determination  of  Clearance  from  Card.  —  Since  it  is  often 
impossible  to  measure  the  clearance  of  an  engine  when  a  test  is 
being  made,  a  rather  rough  approximation  sometimes  is  used. 
Suppose  the  indicator  diagram  is   as  shown  in  Fig.  43.     Two 
points  A  and  B  are  chosen  on  the  compression  curve.     On  these 

two  points  a  rectangle  is 
drawn,  and  the  diagonal  is 
extended  until  it  cuts  the  vol- 
ume axis,  which  is  drawn 
below  the  atmospheric  line  at 
a  distance  equal  to  the  bar- 
ometric pressure.  The  point 
of  intersection  of  this  diagonal 
and  the  volume  axis  establishes  the  origin,  and  the  pressure  axis 
may  be  drawn.  The  distance  c  divided  by  the  length  of  card 
gives  the  ratio  of  clearance.  The  pressure  axis  may  also  be 
located  by  drawing  the  other  diagonal  through  A  and  B  and 
laying  off  DA  equal  to  BE.  If  the  piston  rings  are  not  tight 
in  the  cylinder,  or  if  the  valve  leaks  steam,  this  method  will  not 
give  even  approximately  correct  results. 

90.  Determination  of  the  Mean  Effective  Pressure.  —  It  has 
been  mentioned  that  the  indicator  diagram  shows  the  pressure 
at    all    points    of    the    stroke. 

Since  the  total  pressure  on  the 
piston  times  the  distance  the 
piston  moves  is  the  work  done 
by  the  steam,  we  see  that  the 
indicator  card  is  a  work  dia- 


FIG.  44 


gram. 

In  order  to  calculate  the  work 
done  in  the  cylinder  we  must 
know  the  average  effective  pressure,  or  the  mean  effective  pres- 
sure (m.  e.  p.)  on  the  piston.  In  Fig.  44  when  the  piston  is  at  G 
on  the  forward  stroke,  the  steam  pressure  is  a.  When  the  piston 
is  at  G  on  the  backward  stroke,  the  steam  pressure  is  6.  During 
the  forward  stroke  the  steam  is  working  on  the  piston,  but  on 
the  backward  stroke  the  piston  is  doing  work  on  the  steam. 


THE    STEAM   ENGINE 


91 


Hence  the  effective  pressure  for  the  two  strokes  at  G  is  a  —  b, 
which  is  shown  by  the  dotted  ordinate  G  H. 

To  get  the  average,  or  mean,  of  these  effective  pressures  for  the 
whole  card,  we  may  proceed  as  follows.  Draw  in  the  end-ordi- 
nates  of  the  diagram.  By  means  of  a  scale  or  the  edge  of  a 
ruled  piece  of  paper,  divide  the  length  of  the  card  into  a  number 
of  divisions  of  equal  length.  The  number  of  divisions  should  be 
more  than  eight,  and  need  not  be  more  than  fifteen.  Through 
these  division  points,  draw  in  ordinates  as  shown  by  the  full 
lines.  The  diagram  is  thereby  cut  into  a  number  of  strips  of 
equal  width.  The  average  height  of  each  strip  is  about  equal 
to  the  dotted  ordinate  located  midway  between  the  solid  lines, 


'— *3vm  «/  Jotted  orSin'*t*s' 


FIG.  45 

i.e.  the  average  height  of  the  strip  CEFD  is  GH.  While  this 
may  not  be  true  in  every  case,  the  fact  that  some  parts  of  the 
diagram  are  concave  while  other  parts  are  convex  will  tend  to 
neutralize  the  error  when  the  dotted  ordinates  are  averaged. 
Next  add  the  lengths  of  the  dotted  ordinates.  This  is  best  done 
by  laying  a  strip  of  paper  on  the  diagram  and  marking  the  dif- 
ferent ordinates  directly  on  the  edge  of  the  strip.  Figure  45 
shows  the  strip  of  paper  with  the  dotted  ordinates  added  graph- 
ically. The  average  of  the  ordinates  is  this  sum  divided  by  their 
number.  This  average  height  of  card,  multiplied  by  the  scale 
of  spring,  gives  the  mean  effective  pressure  (m.  e.  p.). 

If  the  m.  e.  p.  of  many  cards  is  to  be  found,  it  may  be  quicker 
to  use  a  polar  planimeter  to  get  the  area  of  the  card.  This  area 
divided  by  the  length  gives  the  mean  height.  This  mean  height 
multiplied  by  the  scale  of  spring  gives  the  m.  e.  p.  This  method 
is  usually  no  more  accurate  than  the  former,  when  the  former 
is  done  with  reasonable  care. 


92  ENGINES   AND   BOILERS 

91.  Indicated  Horsepower.  —  The  mean  effective  pressure 
(m.  e.  p.)  in  pounds  per  square  inch  times  the  area  of  the  piston 
in  square  inches  gives  the  total  average  effective  force  exerted 
by  the  steam  on  the  piston  during  the  forward  and  backward 
stroke,  or  for  one  complete  revolution  of  the  crank.  The  work 
done  by  this  force  is  equal  to  the  force  times  the  distance  the 
piston  moves.  It  must  be  remembered  that  the  effective  force 
on  the  piston  was  obtained  by  taking  the  difference  of  pressures 
during  the  forward  and  backward  strokes.  With  this  in  mind, 
we  see  that  the  work  done  on  the  piston  per  revolution  is  equal  to 
the  m.  e.  p.  times  the  area  of  the  piston,  times  the  length  of  the  stroke. 
If  the  length  of  stroke  is  expressed  in  feet,  the  result  will  be  in 
foot-pounds. 

If  N  denotes  the  revolutions  per  minute  (r.  p.  m.),  L  the  length 
of  stroke  in  feet,  P  the  m.  e.  p.  in  pounds  per  square  inch,  and  A 
the  area  of  piston  in  square  inches,  the  foot-pounds  of  work  done 
per  minute  is  equal  to  PL  A  N.  The  horsepower  of  the  engine  is 
then  PLAAV33000.  Since  this  result  is  obtained  by  means  of 
the  indicator,  it  is  called  the  indicated  horsepower  (i.  hp.). 

If  the  engine  is  double-acting,  the  i.  hp.  for  the  crank-end  is 
found  in  a  similar  manner,  taking  the  m.  e.  p.  from  the  crank- 
end  card  and  using  the  area  of  the  piston  on  the  crank-end,  which 
will  be  less  than  that  for  head-end  because  the  piston  rod  occu- 
pies some  of  the  area  of  the  piston.  For  very  rough  work,  the 
average  of  the  m.  e.  p.  for  the  two  ends  is  sometimes  taken  and 
the  area  of  the  piston  rod  is  neglected;  then  we  have,  approxi- 
mately, 

total  indicated  horsepower  =   QQQQQ  * 

EXAMPLE.  The  head-end  m.  e.  p.  is  42.6  and  the  crank-end  m.  e.  p.  is 
45.1  pounds  per  square  inch  in  a  12"  X 18"  engine  running  at  220  r.  p.  m.  The 
diameter  of  the  piston  rod  is  two  inches.  What  is  the  indicated  horsepower? 

SOLUTION.  The  area  of  a  12"  circle  is  113.1  square  inches  and  of  a  2" 
circle  is  3.1  square  inches.  The  area  of  the  head-end  of  the  piston  is  then 
113.1  square  inches  and  of  the  crank-end  113.1—3.1  =  110.0  square  inches. 
The  stroke  is  18",  or  1.5';  hence  we  have 


and 


.      .       ,.,         42.6X1.5X113.1X220     .__, 
head-end  i.  hp.  =  -  33QQO  -  =48.2  hp., 

45.1X1.5X110.0X220 
crank-end  i.  hp.=  — 33000 —  =49.7  hp. 


whence  the  total  i.  hp.  is  48.2+49.7  =  97.9  hp. 


THE    STEAM   ENGINE 


93 


92.  Brake  Horsepower.  —  It  is  usually  not  difficult  to  take 
cards  from  an  engine  and  to  compute  the  i.  hp.  from  them.  This 
does  not  give  the  horsepower  that  the  engine  is  actually  deliv- 
ering, of  course,  but  that  which  is  developed  in  the  cylinder. 
In  testing  an  engine  that  is  in  actual  use,  it  is  often  impossible 
to  measure  its  actual  output  without  considerable  trouble  and 
expense.  Under  such  conditions  one  must  be  content  with  the 
i.  hp.  If  the  engine  is  not  too  large  and  conditions  will  permit, 
the  actual  power  delivered  by  the  engine  is  often  measured.  If 
it  is  direct-connected  to  an  electric  generator  and  the  losses  in  the 
generator  are  known,  this  is  fairly  simple.  If  not,  some  form  of 
dynamometer  may  be  used.  The  most  common  means  of  meas- 
uring the  delivered  horsepower  is  by  a  brake  on  the  flywheel. 


FIG.  46 

If  a  number  of  tests  are  made,  the  Prony  brake  is  generally  used. 
For  an  occasional  test  a  rope  brake  may  answer  the  purpose. 

Figure  46  shows  a  common  form  of  Prony  brake.  Two  or  more 
bands  of  strap  iron  with  wooden  blocks  fastened  to  them  are 
put  around  the  circumference  of  the  flywheel  or  the  belt  wheel. 
The  tension  in  the  band  is  regulated  by  means  of  a  hand-wheel 
shown  at  the  top  of  the  figure.  Two  arms  are  fastened  to  the 
band  which  at  the  right  end  in  the  figure  carry  a  knife-edge  that 
rests  on  a  block  placed  on  a  platform  scales.  The  flywheel 
rotates  in  the  direction  of  the  arrow,  and  the  friction  between 
the  blocks  and  the  wheel  causes  a  pressure  on  the  scales. 

When  the  engine  is  not  running,  there  will  be  some  pressure 
on  the  scales  due  to  the  weight  of  the  brake-arm.  This  may  be 
determined  by  weighing  the  brake  before  putting  it  on  the  wheel 
and  balancing  it  so  that  its  center  of  gravity  is  determined.  If 


94  ENGINES   AND    BOILERS 

the  weight  of  the  brake  is  w  and  its  center  of  gravity  is  located 
at  a  distance  b  from  the  knife-edge  and  at  a  distance  a  from  the 
center  of  the  wheel,  we  have,  taking  moments  about  the  center 
of  the  wheel, 

wa  =  component  of  weight  on  scales  X  (a +6). 
From  this  we  can  compute  the  effect  of  the  weight  of  the  brake 
arm  on  the  scales.  This  component  of  weight  must  be  subtracted 
from  the  pressure  on  the  scales  when  the  engine  is  running.  If  we 
denote  by  P  the  net  pressure  on  the  scales,  which  is  due  to  the 
friction  between  the  blocks  and  the  wheel  with  the  engine  running, 
we  may  compute  the  power  being  absorbed  by  the  brake  as  fol- 
lows. The  work  absorbed  by  the  brake  per  revolution  is  equal 
to  PX2?rr,  in  which  r  is  the  horizontal  distance  from  the  center 
of  the  wheel  to  the  knife-edge  and  is  known  as  the  radius  of  the 
brake-arm.  If  the  engine  is  running  N  revolutions  per  minute,  the 
work  absorbed  per  minute  is  'ZnrPN,  and  the  horsepower  absorbed 
is  27rrPAV33000.  This  is  called  the  brake  horsepower  (b.  hp.). 
The  radius  of  the  brake  wheel  R  does  not  enter  into  the  compu- 
tation of  the  b.  hp.  To  prevent  the  burning  of  the  wooden  blocks, 
the  wheel  is  kept  cool  by  putting  water  inside  the  rim  which 
evaporates  and  carries  off  the  heat. 

EXAMPLE.  During  the  test  of  an  engine,  the  r.  p.  m.  was  220.  The  pres- 
sure of  the  brake  arm  on  the  scales  was  318  pounds  (Fig.  46).  The  weight  of 
the  entire  brake  is  118  pounds,  and  its  center  of  gravity  (c.  of  g.)  is  6.05 
feet  from  the  knife-edge.  The  radius  of  the  brake-arm,  r,  is  7.16  feet.  Find 
the  brake  horsepower. 

SOLUTION.  If  the  c.  of  g.  of  the  brake  is  6.05  feet  from  the  knife-edge,  it 
is  7.16—6.05  =  1.11  feet  from  the  center  of  the  wheel  when  the  c  of  g.  is  in 
line  between  the  knife-edge  and  the  center  of  the  wheel.  That  part  of  the 
weight  of  the  brake  supported  by  the  scales  is  118X1.11/7.16  =  18.4  pounds. 
The  net  pressure  on  the  scales  then  equals  318  —  18.4  =  299.6  pounds.  The 
brake  horsepower=27rrPAY33000  =  27rX7.16X299.6X220/33000  =  89.7  hp. 

93.  Mechanical  Efficiency.  —  The  ratio  of  the  brake  horse- 
power to  the  indicated  horsepower  is  called  the  mechanical  effi- 
ciency. It  is  usually  expressed  in  the  form  of  a  percentage. 
The  difference  between  the  indicated  horsepower  and  the  brake 
horsepower  is  the  frictional  horsepower. 

EXAMPLE.  If  the  i.  hp.  of  an  engine  is  97.9,  and  the  b.  hp.  is  89.7  hp. 
find  the  mechanical  efficiency  and  the  frictional  horsepower. 

SOLUTION.  The  mechanical  efficiency  =  89.7/97.9  =  .916  =  91.6%.  The 
frictional  horsepower  =  97.9  - 89.7  =  8.2  hp.  Hence  the  frictional  horsepower 
=  8.2/97.9  =  8.4%  of  the  indicated  horsepower. 


THE    STEAM   ENGINE  95 

94.  Thermal  Efficiency.  —  In  general  the  efficiency  of  a  ma- 
chine is  the  ratio  of  the  out-put  to  the  in-put.  In  the  steam 
engine  heat  is  put  in  and  mechanical  work  is  taken  out.  The 
thermal  efficiency  of  the  steam  engine  is  the  ratio  of  the  work  got 
out  to  the  work  equivalent  of  the  heat  put  in.  The  thermal  effi- 
ciency may  be  based  on  either  the  i.  hp.  or  the  b.  hp.  The  latter 
is  called  the  overall  efficiency,  and  is  equal  to  the  thermal  efficiency 
based  on  i.  hp.  times  the  mechanical  efficiency. 

A  common  but  approximate  way  of  stating  the  efficiency  of  a 
steam  engine  is  to  give  the  weight  of  dry  steam  consumed  per 
hour  per  horsepower.  This  may  be  based  on  either  the  i.  hp. 
or  the  b.  hp.  It  is  not  an  exact  way  of  stating  the  efficiency 
because  steam  may  contain  different  amounts  of  heat,  depending 
on  pressures  and  superheat.  Efficiency  may  also  be  expressed 
in  terms  of  B.t.u.  per  minute  per  horsepower. 

In  computing  the  B.t.u.  given  to  the  engine  in  a  unit  of  time, 
proceed  as  follows.  The  weight  of  dry  steam  is  found  by  deduct- 
ing the  weight  of  moisture  in  the  steam  from  the  amount  of 
wet- steam  used.  The  heat  in  this  moisture  is  not  charged  to 
the  engine,  since  it  is  not  possible  for  the  engine  to  extract  work 
from  it.  From  the  steam  tables  find  the  total  heat  in  a  pound 
of  dry  steam,  or  the  total  heat  in  a  pound  of  superheated  steam 
if  superheated  steam  is  used.  From  this  total  heat  per  pound, 
subtract  the  heat  of  the  liquid  at  the  pressure  of  the  exhaust. 
The  reason  for  subtracting  the  heat  of  the  liquid  at  the  pressure 
of  the  exhaust  is  that  although  the  engine  has  used  the  steam, 
the  heat  of  the  liquid  can  be  saved  by  feeding  the  condensed 
steam  back  to  the  boiler,  which  is  often  done.  Whether  it  is 
done  or  not,  it  is  not  fair  to  charge  this  heat  to  the  engine.  Mul- 
tiply the  amount  of  heat  in  a  pound  of  dry  steam  that  is  charged 
to  the  engine  by  the  weight  of  dry  steam  used  in  unit  time.  The 
result  gives  the  B.t.u.  upon  which  the  efficiency  is  computed. 

EXAMPLE.  An  engine  during  a  test  developed  97.9  i.  hp.  and  89.7  b.  hp. 
The  engine  used  3060  pounds  of  97%  quality  steam  per  hour.  The  steam 
pressure  was  125  pounds  gage,  and  the  engine  exhausted  to  the  atmosphere. 
The  barometer  reading  was  29.3  inches.  Find  the  thermal  efficiency  of  the 
engine. 

SOLUTION.  The  weight  of  dry  steam  used  per  hour  is  3060 X. 97  =  2970 
pounds.  125  pounds  gage  pressure  =  125+ 14.4  =  139. 4  pounds  per  square 
inch  absolute.  At  this  pressure,  the  total  heat  in  a  pound  of  steam  is  318. 2 -j- 
872.3  =  1190.5  B.t.u.  The  heat  of  the  liquid  at  the  atmospheric  pressure  is 


96  ENGINES   AND    BOILERS 

179.3,  therefore  the  heat  to  be  charged  to  the  engine  per  pound  of  dry  steam 
used  is  1190.5-179.3  =  1011.2  B.t.u.  The  dry  steam  used  per  hour  per  i.  hp. 
is  2970/97.9  =  30.35  pounds.  The  work  delivered  per  hour  per  horsepower  = 
33000X60  foot-pounds.  The  foot-pounds  of  energy  equivalent  to  the  B.t.u. 
supplied  per  hour  per  horsepower  is  778X1011.2X30.35.  Therefore,  since 
the  efficiency  is  the  out-put  divided  by  the  in-put, 


This  is  based  on  the  i.  hp.  The  dry  steam  used  per  hour  per  b.  hp.  is 
2970/89.7  =  33.1  pounds;  hence  the  thermal  efficiency  based  on  the  b.  hp.  is 

33000X60 
778X1011.2X33.1 

In  this  solution  the  factor  33000X60/778=2545  will  always 
occur  in  the  equation  for  thermal  efficiency.  Hence  the  formula 
may  be  written  in  the  form 

4i         i    a  •  2545 

thermal  efficiency  =  -    .  p  .  —  ,    ,  ,  --  -  -  > 

nX  B.t.u.  per  pound  of  dry  steam 

where  n  is  the  number  of  pounds  of  dry  steam  used  per  hour  per 
horsepower. 

The  thermal  efficiency  of  a  steam  engine  will  seldom,  if  ever, 
exceed  25  per  cent.  This  may  seem  to  be  a  very  low  value,  but 
it  is  impossible  for  the  engine  to  use  a  very  large  part  of  the 
heat  supplied  due  to  the  fact  that  the  exhaust  steam  carries 
with  it  its  heat  of  vaporization.  On  account  of  condensation  of 
steam  in  the  cylinder  and  other  causes  of  heat  loss,  the  efficiency 
of  a  reciprocating  engine  seldom  approaches  the  efficiency  of  an 
ideally  perfect  engine  working  under  the  same  range  of  pressure, 
but  a  well-designed  steam  turbine  of  large  size  may  do  so. 

95.  Cylinder  Condensation.  —  The  largest  single  loss  in  the 
average  engine  is  due  to  what  is  known  as  initial  condensation. 
Since  the  cylinder  walls  are  made  of  iron,  which  is  a  good  con- 
ductor of  heat,  they  naturally  absorb  heat  from  any  hotter 
body  or  substance  placed  in  contact  with  them  and  they  give 
up  heat  to  a  cooler  body.  The  steam  comes  into  the  cyl- 
inder at  a  relatively  high  pressure  and  temperature.  Both  the 
pressure  and  the  temperature  drop  in  the  cylinder,  and  the  steam 
leaves  at  a  relatively  low  pressure  and  temperature.  Since  the 
cylinder  walls  are  exposed  first  to  hot,  and  then  to  cool  steam, 
their  temperature  will  never  be  as  great  as  that  of  the  incoming 
steam,  nor  as  low,  during  operation,  as  that  of  the  outgoing  steam. 

When  the  steam  first  enters  the  cylinder  and  strikes  the  cooler 
walls,  a  part  of  its  heat  will  be  absorbed  by  the  walls.  This 


THE    STEAM   ENGINE  97 

causes  a  partial  condensation  of  the  steam.  Since  the  engine 
operates  by  virtue  of  the  steam  pressure  and  volume,  it  is  readily 
seen  that  a  shrinkage  in  volume  causes  a  loss  of  work,  and  a  lower- 
ing of  efficiency.  By  the  time  the  steam  leaves  the  cylinder,  it 
is  cooler  than  the  cylinder,  and  it  takes  back  some  of  the  heat 
it  gave  to  the  walls,  but  at  too  late  a  time  to  avoid  the  loss  in 
efficiency.  Depending  upon  the  type  of  engine  and  the  condi- 
tions of  operation,  the  condensation  may  continue  until  release 
occurs,  or  re-evaporation  may  start  during  the  expansion  of  the 
steam  between  cut-off  and  release.  By  computing  from  the  in- 
dicator diagram  the  weights  of  steam  at  cut-off  and  at  release, 
we  find  a  net  condensation  during  expansion  if  the  weight  at 
release  is  less  than  at  cut-off,  and  a  net  re-evaporation  if  the 
weight  at  release  is  greater  than  at  cut-off.  The  computation  of 
condensation  or  re-evaporation  during  expansion  is  of  little  value 
since  most  of  the  re-evaporation  occurs  after  release  and  before 
the  steam  leaves  the  exhaust  ports. 

96.   Steam  Accounted  for  by  the  Indicator  Diagram.  —  The 

A.  S.  M.  E.  code  for  testing  steam  engines  calls  for  the  computa- 
tion of  the  steam  accounted  for  by  the  indicator  diagram  at  points 
near  the  cut-off  and  release.  Mark  the  points  of  cut-off  and 
release  and  a  point  on  the 
compression  curve  where  we 
are  sure  the  exhaust  valve  is 
closed,  as  in  Fig.  47.  Find  the 
ratio  of  stroke  at  these  points. 
The  volume  back  of  the  piston 
at  cut-off  is  the  ratio  of  stroke 
at  cut-off  plus  the  ratio  of  pIG  47 

clearance,  shown  by  a,  times 

the  piston  displacement.  Scaling  the  pressure  at  cut-off  from  the 
diagram,  we  may  compute  the  weight  of  dry  steam  back  of  the 
piston  at  cut-off  by  means  of  steam  tables. 

Not  all  of  the  steam  back  of  the  piston  at  cut-off  entered  on 
that  one  stroke  from  admission  to  cut-off,  since  some  of  it  was 
in  the  cylinder  during  compression.  The  amount  that  was  ad- 
mitted is  the  weight  at  cut-off  minus  the  weight  caught  at  com- 
pression. The  weight  of  steam  compressed  may  be  computed 
in  a  manner  similar  to  that  at  cut-off.  The  weight  of  steam 
per  hour  accounted  for  by  the  indicator  diagram  is  then  equal  to 


98 


ENGINES   AND   BOILERS 


FIG.  48 


FIG.  49 


FIG.  50 


FIG.  51 


FIG.  52 


FIG.  53 


FIG.  54 


FIG.  55 


FIG.  56 


FIG.  57 


THE    STEAM   ENGINE  99 

(Wc.o.-TFComp.)XWx60/i.  hp.,  where  Wc.0.  and  TFcomp.  are  the 
weights  of  steam  back  of  the  piston  at  cut-off  and  compression, 
respectively.  The  weight  is  calculated  separately  for  the  head-end 
and  crank-end  of  the  cylinder,  and  the  two  values  added  to  give 
the  total  for  the  engine.  The  weight  accounted  for  at  release  is 
computed  in  the  same  way,  but  the  two  results  usually  differ  slightly 
on  account  of  the  net  condensation  or  the  re-evaporation  during  ex- 
pansion from  cut-off  to  release.  The  weight  of  steam  accounted  for 
by  the  indicator  diagram  will  be  considerably  less  than  the  actual 
amount  used  by  the  engine,  because  of  the  initial  condensation 
of  steam  when  it  first  enters  the  cylinder. 

97.  Valve-setting  from  the  Indicator  Diagram.  —  It  has  been 
mentioned  previously  that  one  of  the  uses  of  the  indicator  is  to 
assist  in  the  setting  of  the  valves  While  the  subject  of  valve- 
setting  will  not  be  discussed  thoroughly  here,  faulty  setting  may 
be  recognized  from  the  appearance  of  the  diagram.  Figure  48 
shows  the  effect  when  the  admission  valve  opens  too  soon. 

Figure  49  shows  the  results  of  late  admission.  Due  to  the 
tardiness  of  the  valve  opening,  the  steam  is  throttled,  and  the 
pressure  for  a  large  part  of  the  stroke  is  lowered  considerably. 

Figure  50  shows  the  effect  of  early  compression.  The  steam 
caught  when  the  exhaust  valve  closes  is  compressed  to  a  pressure 
above  that  in  the  steam  chest,  the  admission  valve  is  lifted  off 
its  seat,  and  some  of  the  steam  escapes  into  the  steam  chest. 

Figure  51  shows  almost  no  compression.  This  would  cause  no 
harm  in  a  very  slow-speed  engine,  but  with  higher  speeds  the 
steam  caught  at  compression  acts  as  a  cushion  and  makes  for 
smooth  running. 

Figure  52  shows  too  early  a  release,  and  Fig.  53,  too  late  a 
release.  Both  cause  a  loss  in  the  area  of  the  diagram. 

Figures  54  and  55  show  unequal  cut-off  in  the  two  ends  of  the 
cylinder.  The  crank-end  is  doing  a  much  larger  proportion  of 
the  work.  The  work  done  by  the  two  ends  should  be  about  equal. 

Figure  56  shows  improper  lubrication  of  the  indicator  piston 
or  the  binding  of  some  part.  A  wavelike  motion  of  the  curve  is 
sometimes  noticed  when  the  diagram  is  taken  from  a  high-speed 
engine,  due  to  the  vibration  of  the  indicator  spring,  but  it  differs 
materially  from  Fig.  56.  In  Fig.  57,  the  indicator  drum  is  strik- 
ing the  stop  on  account  of  improper  adjustment  of  the  length  of 
the  cord  connecting  the  indicator  and  the  reducing  mechanism. 


CHAPTER  VII 
COMMON  TYPES  OF  STEAM   ENGINES 

98.  Slide-valve    Engine.  —  Where   simplicity   and   reliability 
are  of  more  importance  than  high  efficiency,  the  slide-valve  en- 
gine is  used.     The  simplest  type  of  slide-valve  was  shown  in  Fig. 
39,  and  the  principles  of  its  operation  were  explained  to  some 
extent  in  the  previous  chapter.     There  are  many  varieties  of 
slide-valves,  the  more  common  of  which  will  be  described  later. 
Most  of  the  smaller  stationary  engines  in  use  are  equipped  with 
the  slide-valve,  and  all  American  locomotive  engines  are  of  this 
type. 

While  the  plain  slide-valve  is  very  simple,  it  has  certain  defects. 
One  of  these  is  the  impossibility  of  obtaining  the  proper  steam 
distribution  at  all  loads,  i.e.  of  making  the  events  of  stroke  occur 
at  the  proper  place  to  give  the  highest  efficiency  at  light  load 
and  also  at  heavy  load.  Various  modifications  and  improvements 
have  been  made  on  the  slide-valve  to  remedy  this  defect,  the 
chief  one  of  which  is  to  place  a  second  slide-valve  on  top  of  the 
main  valve,  and  to  control  cut-off  by  a  rider. 

Another  defect  of  the  plain  slide-valve  is  the  slowness  with 
which  the  steam  ports  open  up  and  close  at  some  loads,  which 
cause  what  is  known  as  wire-drawing.  This  is  simply  a  throttling 
of  the  steam  by  the  valve  as  it  enters  the  cylinder.  This  throt- 
tling usually  causes  a  lowering  of  the  efficiency  of  the  engine. 

99.  The  Corliss  Engine.  —  By  far  the  most  common  type  of 
high-grade  reciprocating  stationary  steam  engine  in  this  country 
is  the  Corliss  engine.     The  name  comes  from  its  inventor  and 
first  producer,  GEORGE  CORLISS,  an  engineer  and  engine  builder  of 
Providence,  R.  I.     There  are  two  distinguishing  features  of  this 
engine.     The  first  of  these  is  the  oscillating  cylindrical  valve. 
The  second  is  the  means  for  disengaging  the  valve  from  the  mech- 
anism that  drives  it,  and  the  quick  closing  of  the  valve  after  its 
disengagement.     To   understand   the    Corliss   valve   mechanism 
thoroughly,  it  is  necessary  to  make  a  rather  thorough  analysis  of 
its  motion.     We  shall  not  do  this  in  this  chapter.     The  general 
principle  of  operation  of  the  gear  is  fairly  simple,  however. 

Figure  58  shows  a  typical  Corliss  engine  cylinder.  The  right 
end  is  cut  in  section  to  show  the  construction  of  the  valves  and 

100 


COMMON   TYPES   OF    STEAM   ENGINES 


101 


their  locations  relative  to  the  cylinder.  The  left  end  shows  an 
ordinary  form  of  the  mechanism  that  moves  the  steam  valves  and 
the  exhaust  valves.  An  eccentric  on  the  shaft  is  connected  to 
the  hook  rod  which  operates  the  valves  through  an  eccentric  rod 
and  rocker  arm.  This  gives  the  hook  rod  a  horizontal  recipro- 
cating motion  that  is  nearly  harmonic.  The  hook  rod  is  attached 


-  3 fa  am  fff>e 


FIG.  58 


to  a  wrist  plate  which  is  pivoted  to  the  cylinder  at  C.     The  wrist 
plate  is  thereby  given  an  oscillating  motion. 

Four  rods  are  attached  to  the  wrist  plate.  The  two  upper 
rods  are  the  steam  rods,  which  transmit  the  motion  to  the  two 
steam  valves,  and  the  lower  or  exhaust  rods  drive  the  two  exhaust 
valves.  The  steam  rods  are  attached  to  bell  cranks  or  double 
arms  that  are  pivoted  on  the  valve  spindle  but  are  not  attached 
to  it.  It  is  thus  possible  for  the  wrist  plate  and  the  bell  crank 


102  ENGINES   AND    BOILERS 

to  move  without  affecting  the  valve  in  any  way.  The  cylindrical 
steam  valve  has  a  spindle  which  extends  out  of  the  steam  chest. 
To  the  outer  end  of  this  spindle  there  is  keyed  a  steam  arm. 
Any  motion  of  this  arm  causes  the  valve  to  move.  A  block  is 
attached  to  the  back  side  of  the  steam  arm,  and  a  hook  which  is 
carried  by  the  upper  arm  of  the  bell  crank  catches  over  it.  As 
the  steam  rod  moves  to  the  right,  the  steam  arm  is  picked  up 
and  the  valve  is  turned.  After  having  lifted  the  steam  arm  a 
certain  distance,  the  hook  is  made  to  disengage  -with  the  block 
and  the  steam  arm  is  released. 

The  steam  arm  is  connected  to  the  piston  of  a  dash-pot  by  a 
dash-rod.  As  the  steam  arm  is  raised,  a  partial  vacuum  is 
formed  in  the  dash-pot.  When  the  steam  arm  is  released  from 
the  hook,  it  is  suddenly  pulled  downward  by  the  vacuum  in  the 
dash-pot.  As  the  steam  arm  is  lifted,  the  valve  opens  and  admits 
steam  to  the  cylinder.  When  it  is  pulled  down,  the  valve  is 
closed  suddenly,  giving  a  quick  cut-off.  The  time  at  which  the 
hook  is  made  to  release  the  steam  arm  is  controlled  by  a  cam 
whose  position  is  regulated  by  the  governor.  This  cam  engages 
with  the  tail  of  the  hook  and  causes  the  disengagement.  At  light 
loads,  the  trip  occurs  soon  and  an  early  cut-off  is  given,  and  the 
cut-off  is  retarded  as  the  load  of  the  engine  is  increased. 

Figure  59  shows  the  trip  mechanism  on  a  larger  scale.  DAC 
is  the  bell  crank.  As  the  point  D  moves  backward  and  forward 
in  a  nearly  horizontal  direction,  the  point  C  moves  up  and  down 
in  a  nearly  vertical  direction.  The  pin  C  carries  the  steam  hook. 
The  tail  of  the  hook  engages  with  the  knock-off  cam,  and  its  jaw 
engages  with  the  block  attached  to  the  steam  arm  at  B.  For 
any  one  load  the  knock-off  cam  is  stationary,  and  as  C  goes  up, 
the  tail  of  the  hook  is  pushed  away  from  A  by  the  cam,  which 
causes  the  latch  to  disengage  with  the  block  B.  When  there  is 
a  heavy  load  on  the  engine,  the  governor  rod  moves  to  the  left, 
which  raises  the  knock-off  cam  and  makes  the  trip  come  later, 
giving  a  long  cut-off.  There  is  also  a  safety  cam,  shown  in  Fig. 
59.  If  the  governor  fails  to  rotate,  the  safety  cam  comes  into 
contact  with  the  tail  of  the  hook  and  prevents  the  picking  up  of 
the  steam  arm  and  therefore  causes  a  failure  to  admit  steam  to 
the  cylinder. 

There  is  no  disengagement  between  the  exhaust  arm  and  the 
exhaust  valve,  so  that  the  events  of  release  and  compression  occur 


COMMON    TYPES    OF    STEAM   ENGINES 


103 


at  the  same  ratio  of  the  stroke  for  all  loads.     Both  the  steam 
valve  and  the  exhaust  valve  of  Fig.  58  are  double-ported,  which 
gives  twice  the  opening  for  the  passage  of  steam  with  the  same 
valve  movement  as  with  the  single-ported  type. 
From  the  description  just  given  it  is  seen  that  cut-off  is  inde- 


To  forer/ior 


FIG.  59 


pendent  of  the  other  events,  and  that  the  steam  valve  closes 
quickly,  thereby  preventing  wire-drawing  at  closing.  If  a  care- 
ful analysis  of  the  motion  is  made,  it  will  be  seen  that  the  steam 
valve  opens  the  port  nearly  as  widely  at  light  loads  as  at  full 
loads.  The  force  necessary  to  operate  the  valve  mechanism  is 
not  large  and  the  work  done  in  moving  the  valves  is  a  very  small 
part  of  the  total  output  of  the  engine. 

100.  The  Four-valve  Engine.  —  With  a  single  slide-valve, 
the  changing  of  one  event  necessitates  the  changing  of  all  the 
others.  To  avoid  this  difficulty,  engines  are  often  made  that 


104 


ENGINES   AND    BOILERS 


have  four  valves,  a  steam  valve  for  each  end  of  the  cylinder,  and 
an  exhaust  valve  on  each  end.  The  exhaust  valves  are  driven 
by  a  fixed  eccentric,  so  that  the  release  and  compression  are  the 
same  at  all  loads.  The  steam  valves  are  controlled  by  the  gov- 
ernor; hence  cut-off  and  admission  will  vary  for  different  loads. 


f  FIG.  60 

Figures  60  and  61  show  one  style  of  four- valve  engine.  This  par- 
ticular engine  has  oscillating  cylindrical  valves  similar  to  those 
shown  in  Fig.  58  for  the  Corliss  engine.  Some  makers,  probably 
to  make  use  of  the  enviable  reputation  of  the  Corliss  engine,  call 
this  type  a  non-releasing  Corliss.  This  engine  lacks,  however, 
the  distinct  advantage  of  the  trip  found  in  the  true  Corliss  type. 


FIG.  61 


101.  The  Compound  Engine.  —  When  all  the  expansion  of 
steam  takes  place  in  one  cylinder,  we  have  what  is  known  as  a 
simple  engine.  If  the  steam  passes  through  two  successive  cyl- 
inders, the  engine  is  said  to  be  a  compound  engine.  If  there  are 
three  successive  cylinders,  it  is  called  a  triple-expansion  engine. 


COMMON   TYPES   OF   STEAM   ENGINES  105 

If  there  are  four  successive  cylinders,  it  is  called  a  quadruple- 
expansion  engine,  etc.  An  engine  may  have  two  cylinders  and 
not  be  compound,  i.e.  it  may  be  a  twin-cylinder  engine,  in  which 
half  the  steam  passes  through  one  cylinder  and  half  through  the 
other.  Likewise  a  compound  engine  may  have  three  cylinders, 
all  the  steam  passing  through  one  high-pressure  cylinder,  and  then 
dividing,  half  passing  through  each  of  the  two  low-pressure 
cylinders. 

The  purpose  of  compounding  is  to  reduce  the  initial  conden- 
sation. It  does  not  necessarily  follow  that  there  is  a  greater 
ratio  of  expansion  of  steam  in  a  compound  engine  than  in  a 
simple  engine,  since  that  depends  upon  the  point  of  cut-off.  We 
have  seen  that  initial  condensation  is  caused  by  the  range  in 
temperature  within  the  cylinder.  The  temperature  range  is  less 
in  each  cylinder  of  a  compound  engine  than  in  the  single  cylinder 
of  a  simple  engine  of  the  same  capacity.  Since  the  amount  of 
condensation  does  not  vary  directly  as  the  total  temperature 
range,  there  may  be  considerably  less  total  condensation  if  the 
steam  is  passed  through  two  successive  cylinders  than  if  all  the 
expansion  occurred  in  one  cylinder. 

Several  years  ago  the  idea  of  compounding  was  very  popular 
and  was  carried  to  the  extreme.  Many  triple-expansion  engines, 
and  some  quadruple-expansion  engines,  were  built.  Experience 
proved,  however,  that  there  was  a  practical  limit  to  which  the 
idea  might  be  carried.  Now  stationary  engines  are  seldom  built 
with  more  than  two  pressure-stages,  except  in  direct -acting  pumps. 
In  the  marine  service,  the  triple-expansion  engine  is  still  popular, 
partly  for  the  reason  that  it  is  desirable  to  have  three  cranks  on 
the  same  shaft  to  give  a  greater  uniformity  of  torque  on  the  pro- 
peller shaft,  and  partly  on  account  of  the  uniformity  of  load  on 
marine  engines.  Of  the  many  types  of  compound  engines  that 
have  been  built,  only  two  are  in  common  use  in  land  service  at 
present.  We  shall  proceed  to  consider  these. 

102.  The  Tandem-compound  Engine.  —  In  the  tandem-com- 
pound engine,  the  pistons  of  the  two  cylinders  are  placed  on  the 
same  piston  rod,  as  shown  in  Fig.  62.  The  cylinder  to  the  left 
is  the  high-pressure  cylinder,  and  the  one  to  the  right  is  the  low- 
pressure  cylinder.  The  steam  ports  of  the  high-pressure  cylinder 
are  at  a  and  6;  the  exhaust  ports  of  the  low-pressure  cylinder  are 


106 


ENGINES   AND    BOILERS 


at  c  and  d.  The  pistons,  in  Fig.  62,  are  shown  moving  to  the 
right.  Steam  is  entering  the  high-pressure  cylinder  through  a, 
and  leaving  it  through  /.  The  exhaust  steam  from  the  crank  end 
of  the  high-pressure  cylinder  passes  to  the  head  end  of  the  low- 
pressure  cylinder  either  directly  or  through  a  stationary  vessel 
called  a  receiver,  i.e.  the  back  pressure  on  the  piston  A  is  the 
forward  pressure  on  the  piston  B.  On  the  return  stroke,  steam 


FIG.  62 

enters  the  port  b  and  the  exhaust  from  the  high-pressure  cylinder 
leaves  through  e  and  enters  the  low-pressure  cylinder  at  h  either 
directly  or  through  the  receiver.  With  the  tandem  arrangement 
only  one  cross-head,  connecting  rod,  crank,  and  frame  are  needed. 
In  locomotive  work,  the  Baldwin  or  Vauclain  compound  engine 
is  sometimes  seen.  In  this  engine,  the  cylinders  are  placed  side 
by  side,  and  both  piston  rods  attach  to  the  same  cross-head. 
The  method  of  steam  distribution  is  similar  to  that  of  the  tan- 
dem type.  Figures  63  and  64  show  the  high-pressure  and  low- 


FIG.  63 


FIG.  64 


FIG.  65 


pressure  indicator  diagrams,  as  taken  from  a  Baldwin  compound 
engine.  The  high-pressure  card  comes  from  the  crank  end  of  the 
high-pressure  cylinder  and  the  low-pressure  card  from  the  head 
end  of  the  low-pressure  cylinder.  These  cards  were  taken  with 
springs  of  different  scales.  Figure  65  shows  the  same  diagrams  when 
drawn  to  the  same  scale  of  pressure.  It  is  noticed  that  the  back- 
pressure line  of  the  high-pressure  card  parallels  the  admission 
line  of  the  low-pressure  card  from  the  left  end  up  to  the  point 
of  cut-off  in  the  low-pressure  cylinder.  The  reason  for  this  is 


COMMON   TYPES    OF    STEAM   ENGINES 


107 


obvious,  since  the  exhaust  from  the  high-pressure  cylinder  passes 
directly  to  the  low-pressure  cylinder.  When  the  admission 
valve  of  the  low-pressure  cylinder  closes,  compression  must 
necessarily  start  in  the  high-pressure  cylinder.  Since  it  is  neces- 
sary, with  such  a  high  pressure  at  exhaust  as  exists  in  the  high- 
pressure  cylinder,  to  have  the  compression  occur  late,  it  follows 
that  cut-off  must  come  very  late  in  the  low-pressure  cylinder. 
This  is  not  an  ideal  condition,  but  it  is  necessary  if  no  receiver 
is  placed  between  the  two  cylinders.  If  a  receiver  were  placed 
between  the  two  cylinders  so  that  it  could  act  as  a  reservoir  into 
which  to  discharge,  and  from  which  to  draw  steam,  it  would  not 
be  necessary  to  have  the  preceding  relation  between  compression 
and  cut-off. 

103.  The  Cross-compound  Engine.  —  In  the  cross-compound 
engine  (Fig.  66),  each  cylinder  has  its  own  cross-head,  connect- 


" 


FIG.  66 

ing  rod,  crank,  and  frame.  The  cranks  are  usually  spaced  90°  apart. 
A  by-pass  is  arranged  so  that  the  engine  can  be  started  even  if 
the  high-pressure  cylinder  stops  on  dead  center,  by  admitting 
steam  directly  to  the  low-pressure  cylinder. 

Each  cylinder  has  its  own  valve  mechanism,  and  the  exhaust 
from  the  high-pressure  cylinder  passes  into  a  receiver  from  which 
the  low-pressure  cylinder  takes  its  steam.  This  arrangement  per- 
mits a  better  steam  distribution  than  that  used  in  the  tandem 


108 


ENGINES   AND    BOILERS 


type  without  a  receiver.  If  the  receiver  is  quite  large,  the  back 
pressure  in  the  high-pressure  cylinder  during  exhaust  will  be 
nearly  constant.  Figures  67  and  68  show  the  indicator  diagrams 
from  a  cross-compound  engine.  It  should  be  noticed  that  the 
engine  exhausts  into  a  condenser. 

104.  Cylinder  Ratio.  —  The  cylinder  ratio  of  a  compound  engine 
is  the  ratio  between  the  piston  displacements  of  the  low-pressure  and 
the  high-pressure  cylinders.  While  it  is  not  essential  that  the  length 
of  stroke  be  the  same  for  both  high-pressure  and  low-pressure 

cylinders  of  a  cross- 
compound  engine,  they 
are  made  so.  The  cyl- 
inder ratio  is  then  the 
ratio  of  the  squares  of 
the  diameters  of  the 
low-pressure  and  high- 
pressure  cylinders. 

105.  The  Combined 
Indicator    Diagram.  — 

The  combined  diagram 
is  constructed  by  plot- 
ting both  cards  to  the 
same  scale  of  pressure 
and  volume.  Usually 
we  do  not  change  the 
low-pressure  diagram 
but  change  the  scale  of 
the  high-pressure  card 
conform  to  it.  Figure 
69  shows  the  combina- 
tion of  diagrams  of  Figs. 
67  and  68.  The  low- 
pressure  diagram  is 
identical  with  Fig.  68,  while  the  length  of  the  high-pressure  dia- 
gram equals  the  length  of  the  low-pressure  diagram  divided  by 
the  cylinder  ratio.  The  high-pressure  diagram  is  placed  to  the 
right  of  the  pressure  axis  its  clearance  distance,  i.e.  its  distance 
from  the  axis  equals  the  ratio  of  the  high-pressure  clearance 
times  the  new  length  of  the  high-pressure  diagram. 


Condenser  Pressure. 

FIG.  69 


COMMON    TYPES    OF    STEAM    ENGINES  109 

106.  Diagram  Factor. —  The  definition  of  the  diagram  factor 
as  given  in  the  1915  edition  of  the  A.  S.  M.  E.  Power  Test  Code  is 
as  follows : 

The  diagram  factor  is  the  proportion  borne  by  the  mean 
effective  pressure  measured  from  the  actual  diagram  to  that 
of  a  hypothetical  diagram  which  represents  the  maximum  power 
obtainable  from  the  steam  accounted  for  by  the  actual  diagram 
at  the  point  of  cut-off;  assuming  first,  that  the  engine  has  no 
clearance;  second,  that  there  are  no  losses  through  wire- 
drawing the  steam  either  during  admission  or  release;  third, 
that  the  expansion  line  is  a  hyperbolic  curve;  and,  fourth, 
that  the  initial  pressure  is  that  of  the  boiler,  and  the  back 
pressure  that  of  the  atmosphere  for  a  non-condensing  engine, 
and  of  the  condenser  for  a  condensing  engine. 
To  determine  the  steam  accounted  for  by  the  actual  diagram 
at  the  point  of  cut-off,  draw  hyperbolic  curves  through  the  point 
of  compression  P  and  the  point  of  cut-off  0  (Fig.  70)  until  they 


FIG.  70 

cut  the  boiler-pressure  line  at  R  and  S.  The  length  of  RS 
is  the  length  of  the  admission  line  for  the  hypothetical  diagram, 
FA  in  Fig.  42,  drawn  to  proper  scale.  The  hypothetical  diagram 
is  drawn  as  in  Fig.  42  except  that  the  boiler  pressure  is  taken  as  the 
initial  pressure,  release  comes  at  the  end  of  the  stroke,  the  back 
pressure  is  the  atmospheric  pressure  (condensing  pressure  in  a  con- 
densing engine),  there  is  no  compression,  and  there  is  no  clear- 
ance. The  hypothetical  diagram  for  the  combined  diagrams  of 
Fig.  69  is  shown  dotted.  Since  we  assume  there  is  no  clearance, 
the  length  of  the  hypothetical  diagram  is  equal  to  that  of  the 
low-pressure  card.  The  distance  RS  at  boiler  pressure  is  deter- 
mined from  the  high-pressure  diagram,  as  in  Fig.  70.  From  S 
to  T  construct  a  hyperbolic  curve,  using  the  origin  0',  and  not  0. 
Release  is  at  the  end  of  the  stroke  and  the  back-pressure  line  is 
at  the  condenser  pressure. 


*-*'*'•  /}b 
i<$  CX^Trar 


110  ENGINES   AND    BOILERS 

In  Fig.  69,  the  mean  effective  pressure  of  the  combined  dia- 
grams and  of  the  hypothetical  diagram  are  in  the  same  ratio  as 
the  areas  of  the  combined  and  hypothetical  diagrams,  because 
they  are  of  the  same  length.  To  find  the  diagram  factor  of  the 
combined  cards,  divide  their  area  by  the  area  of  the  hypothetical 
diagram. 

107.  Ratio  of   Expansion.  —  The  A.   S.   M.   E.   Power  Test 
Code  gives  the  following  rule: 

To  find  the  percentage  of  cut-off,  or  what  may  best  be  termed 
the  commercial  cut-off,  the  following  rule  should  be  observed: 

Through  the  point  of  maximum  pressure  during  admission 
draw  a  line  parallel  to  the  atmospheric  line.  Through  a  point 
on  the  expansion  line  where  the  cut-off  is  complete,  draw  a 
hyperbolic  curve.  The  intersection  of  these  two  lines  is  the 
point  of  commercial  cut-off,  and  the  proportion  of  cut-off  is 
found  by  dividing  the  length  measured  up  to  this  point  by 

e  total  length> 

To  find  the  ratio  of  expansion,  divide  the  volume  correspond- 

bg  to  the  piston  displacement,  including  clearance,  by  the 
olume  of  the  steam  at  the  commercial  cut-off,  including  clear- 

ance. 

In  a  multiple-expansion  engine  the  ratio  of  expansion  is  found 

by  dividing  the  volume  of  a  low-pf  essure  cylinder,  including 

clearance,  by  the  volume  of  the  high-pressure  cylinder  at  the 

commercial  cut-off,  including  clearance. 

108.  The  Unaflow  Engine.  —  The  unaflow,  or  uniflow,  engine 
is  shown  diagrammatically  in  Fig.  71.     There  is  an  admission 
valve  at  each  end  of  the  cylinder.     The  exhaust  steam  escapes 
through  a  port  located  around  the  circumference  of  the  cylinder 
midway  between  the  two  ends.     The  piston,  which  is  longer  than 
in  most  engines,  itself  uncovers  the  exhaust  port  at  about  90 
per  cent  of  the  stroke.     Compression  must  start  when  the  piston 
is  at  the  same  place  on  the  return  stroke.     Under  non-condensing 
conditions  this  would  give  a  very  excessive  compression  pressure; 
hence  the  engine  normally  is  run  condensing,  under  which  con- 
ditions the   compression    pressure    is    moderate.     The    thermal 
efficiency  is  about  the  same  as  that  of  a  compound  engine.     The 
gain  in  efficiency  over  the  ordinary  double  flow  engine  is  due 
to   the   reduction    of    initial    condensation.     The    condensation 


COMMON    TYPES    OF    STEAM   ENGINES 


111 


is  reduced  with  the  unaflow  principle  because  the  ends  of 
the  cylinder  are  kept  hotter  than  the  central  portion.  High- 
pressure  steam  never  comes  in  contact  with  the  central  part  of 
the  cylinder  and  the  flow  of  steam  is  from  the  ends  toward  the 
middle.  The  exhaust  steam  passing  out  through  the  central  port 
does  not  cool  the  walls  as  much  as  it  would  if  it  flowed  back  to 


the  ends  of  the  cylinder  upon  leaving.  In  actual  engines,  pro- 
vision must  be  made  for  relieving  the  excessive  compression  pres- 
sure, should  the  vacuum  break.  This  is  done  by  a  relief  valve 
that  adds  to  the  clearance  or  allows  the  compressed  steam  to 
re-enter  the  steam  chest,  or  by  adding  an  auxiliary  exhaust  port 
nearer  the  end  of  the  cylinder,  which  is  opened  automatically 
when  the  vacuum  fails. 


CHAPTER  VIII 
VALVES 

109.  Introduction.  —  From  our  previous  study  of  the  steam 
engine  we  have  learned  that  the  purpose  of  the  valve  is  to  admit 
steam  to  the  cylinder,  and  to  release  steam  from  it.     The  time 
at  which  the  events  occur  must  be  such  that  the  engine  is  capable 
of  doing  the  work  required,  and  that  it  may  have  as  high  an 
efficiency  as  possible  under  the  conditions  of  operation.     An  en- 
gine may  run  with  the  valves  improperly  set  or  designed,  but 
more  steam  will  be  used  than  if  the  valves  functioned  properly. 

110.  The  D  Slide-valve.  —  The  engine  of   Fig.  39  has  what 
is  commonly  known  as  a  D  slide-valve.     The  valve  slides  back 
and  forth  on  its  seat,  alternately  opening  and  closing  the  ports. 
An  eccentric  on  the  shaft  drives  the  valve.     The  eccentric  rod 
is  usually  quite  long  in  comparison  with  the  throw  of  the  eccen- 
tric, so  that  the  valve  may  be  considered  to  have  the  same  motion 
as  the  horizontal  component  of  the  eccentric. 

In  Fig.  72a,  the  valve  is  shown  in  mid-position;  consequently 
the  eccentric  will  either  be  directly  above  or  directly  below  the 
center  of  the  shaft.  In  mid-position,  the  valve  laps  over  the 
edges  of  the  port;  the  amount  it  extends  over  on  the  steam  side 
is  called  the  steam  lap,  and  on  the  exhaust  side,  the  exhaust  lap. 

At  the  right  side  of  Fig.  72a  is  shown  the  relative  position  of 
the  eccentric  and  the  crank.  The  eccentric  leads  the  crank  by 
angle  8.  This  angle  8  will  be  the  same  at  all  times  during  the 
revolution.  In  Fig.  726,  the  crank  is  on  head-end  dead  center, 
and  the  valve  is  uncovering  the  head-end  port  a  small  amount. 
The  amount  the  port  is  open  when  the  crank  is  on  dead  center 
is  called  the  lead.  It  is  measured  in  inches.  The  valve  in  Fig. 
726  is  to  the  right  of  its  mid-position  by  an  amount  equal  to  the 
steam  lap  plus  the  lead.  In  moving  from  the  position  shown 
in  Fig.  72a  to  that  in  Fig.  726,  the  eccentric  has  moved  a  hori- 
zontal distance  equal  to  the  steam  lap  plus  the  lead,  and  has  turned 
through  a  certain  angle  which  is  called  the  angle  of  advance. 
It  is  evident  that  the  eccentric  is  ahead  of  the  crank  by  an  angle 
of  90°  plus  the  angle  of  advance.  It  is  customary  to  speak  of 
the  angle  of  advance  and -not  of  the  whole  angle  6. 

Figure  72c  shows  the  valve  at  head-end  admission.  The  valve 

112 


VALVES 


113 


FIG.  72 


114  ENGINES   AND   BOILERS 

is  on  the  point  of  opening  the  head-end  port  for  the  admission 
of  steam,  and  is  traveling  to  the  right.  In  the  admission  posi- 
tion it  is  to  the  right  of  its  mid-position  by  a  distance  equal  to 
the  steam  lap. 

Likewise  the  eccentric  will  be  to  the  right  of  its  mid-position 
by  a  horizontal  distance  equal  to  the  steam  lap.  The  crank 
is  back  of  the  eccentric  by  the  angle  6  and  is  seen  to  be 
approaching  its  head-end  dead-center  position.  If  there  is 
any  lead,  the  crank  will  never  be  quite  up  to  its  dead-center 
position  at  admission. 

At  head-end  cut-off,  Fig.  72d,  the  valve  is  to  the  right  of  its 
mid-position  by  a  distance  equal  to  the  steam  lap  and  its  direc- 
tion of  motion  is  to  the  left.  The  position  is  the  same  as  for 
admission,  but  it  is  going  in  the  opposite  direction.  The  eccen- 
tric is  now  below  the  center  line  of  the  shaft. 

Figures  72e  and  72/  show  the  relative  positions  at  head-end 
release  and  compression.  At  both  of  these  events,  the  valve  is 
at  the  left  of  mid-position  by  an  amount  equal  to  the  exhaust- 
lap  distance.  It  is  moving  to  the  left  at  release  and  to  the  right 
at  compression. 

In  all  the  six  diagrams  of  Fig.  72,  the  positions  of  the  piston 
and  its  direction  of  motion  are  shown.  The  cylinder  section  in 
each  diagram  is  in  a  horizontal  plane  and  therefore  is  at  an  angle 
of  90°  from  the  diagram  showing  crank  and  eccentric  positions 

For  an  understanding  of  the  slide-valve  and  the  analyses  to  follow, 
it  is  essential  that  the  student  have  a  precise  conception  of  the  rel- 
ative positions  of  the  valve  on  the  seat,  the  eccentric  and  the  crank 
relative  to  the  center  positions,  and  the  position  of  the  piston  in  the 
cylinder. 

111.  Relative  Motion  of  Crank  and  Piston.  —  Since  the  con- 
necting rod  is  not  very  long  compared  with  the  crank  arm,  we 
cannot  assume  that  the  horizontal  movement  of  the  crank  is  the 
same  as  the  piston  movement.  This  is  clearly  seen  from  the 
diagram  in  Fig.  73.  As  the  crank  moves  from  A  to  C,  the  cross- 
head  moves  from  E  to  D.  That  part  of  the  stroke  completed 
by  the  cross-head  is  a'.  It  is  evident  that  a'  is  considerably 
larger  than  the  horizontal  movement  of  the  crank  in  going  from 
A  to  C. 

In  our  analysis  of  valve  motions,  it  is  not  customary  to  draw 


VALVES  115 

in  the  cross-head  D  to  find  the  proportion  of  the  stroke  at  dif- 
ferent crank  positions,  but  the  following  scheme  is  used.  The 
horizontal  diameter  of  the  crank  circle  AB  is  extended  to  the 
left.  With  the  length  of  connecting  rod  DC  as  a  radius  at  the 
desired  scale,  and  with  D  as  a  center,  strike  the  arc  shown  by 
the  dotted  line  CG.  This  gives  the  distance  AG  =  a,  on  the  diame- 
ter of  the  crank  circle,  that  is  equal  to  ED  =  a',  the  movement  of 


FlG.  73 

the  piston  or  cross-head  from  E  to  D.  In  the  valve  analysis  it 
is  not  necessary  to  draw  the  crank  circle  to  any  particular  scale 
if  we  keep  the  proper  ratio  between  the  lengths  of  the  crank  and 
the  connecting  rod.  This  ratio  usually  is  expressed  as  R/L. 
No  matter  what  the  scale  of  the  crank  circle  may  be,  the  ratio  of 
a  to  the  length  of  stroke  will  be  constant. 

112.  Valve  Diagrams.  —  Many  diagrams  have  been  used  to 
show  graphically  the  relation  of  the  movement  of  the  valve  to 
the  movement  of  the  piston,  or  of  the  relative  movements  of 
eccentric  and  crank.     Only  those  most  commonly  used  in  this 
country  will  be  explained  here,  i.e.  the   valve  ellipse,  the    Bil- 
gram  diagram,  and  the  Zeuner  diagram. 

113.  The  Valve  Ellipse. — In  this  diagram  the  system  of  rec- 
tangular coordinates  is  used.     The  valve  displacement  is  plotted 
vertically  and  the  piston  displacements  are  plotted  horizontally. 
On  the  left  of  Fig.  74  is  shown  a  crank  and  eccentric.   The  eccen- 
tric is  ahead  of  the  crank  by  90°  plus  a. 

With  the  crank  at  C,  the  piston  is  at  a  distance  x  from  the  cen- 
ter of  the  stroke.  At  the  same  time,  the  valve  is  at  a  distance  y 
from  its  mid-position.  If  we  plot  x  against  y,  we  get  a  point  G. 


116 


ENGINES   AND   BOILERS 


The  coordinate  axes  are  the  horizontal  and  vertical  diameters 
of  the  crank  circle. 

On  the  right  of  Fig.  74  this  same  operation  is  carried  out  for 
twelve  crank  positions  with  their  corresponding  eccentric  posi- 
tions. The  crank  positions  are  denoted  by  Ci,  €2,  C%,  etc.,  and 
the  corresponding  eccentric  positions  by  E\,  E^  E%,  etc.  Plotting 
the  displacements,  we  get  the  points  1,  2,  3,  etc.  Connecting  the 
points  thus  found  by  a  smooth  curve,  we  get  what  is  known  as 
a  valve  ellipse.  It  is  evident  from  Fig.  74  that  this  is  not  a  true 


fro  re/ '  ~  -  -^ 


FIG.  74 

ellipse.  It  would  have  been  but  for  the  distortion  due  to  the 
short  length  of  the  connecting  rod. 

The  upper  half  of  the  ellipse  ABD  represents  the  valve  move- 
ment to  the  right  of  its  mid-position.  The  lower  half,  DFA, 
represents  the  valve  movement  to  the  left  of  its  mid-position. 
From  A  to  B  it  gives  the  movement  from  mid-position  to  ex- 
treme right;  from  B  to  D,  from  the  extreme  right  to  mid-posi- 
tion; from  D  to  F  from  mid-position  to  extreme  left;  and  F  to 
A,  from  the  extreme  left  to  mid-position. 

The  valve  ellipse  of  Fig.  74  is  reproduced  in  Fig.  75.  Four 
horizontal  lines  are  drawn  through  the  ellipse.  The  head-end 
steam  lap  is  the  distance  from  the  top  line  to  the  horizontal  axis. 
The  crank-end  steam  lap  is  the  distance  from  the  bottom  line 
to  the  axis.  The  head-end  exhaust-lap  line  is  drawn  below  the 
axis  and  the  crank-end  exhaust-lap  line  above.  When  the  valve 
has  moved  to  the  right  a  distance  equal  to  the  head-end  steam 
lap,  head-end  admission  takes  place.  Admission  is  shown  by  the 


VALVES 


117 


point  H  on  the  ellipse,  and  the  crank  position  corresponding  to 
H  is  determined  by  projecting  vertically  from  H  to  the  axis. 

As  the  valve  moves  from  its  extreme  right  position  back  to 
mid-position,  head-end  cut-off  takes  place.  This  is  shown  by  the 
point  /  on  the  ellipse.  The  crank  position  corresponding  to  I  is 
found  by  projecting  down  from  I  to  the  axis  and  striking  an 
arc  upward  from  this  point  to  the  crank  circle.  The  radius  of 
the  arc  is  the  length  of  the  connecting  rod.  In  like  manner,  the 


A 

•fxg/w 

Ce.mof.ft.0. 


FIG.  75 


crank  position  at  head-end  release  is  found  from  the  intersection 
of  the  head-end  exhaust-lap  line  with  the  ellipse  at  J. 

The  head-end  compression  point  is  at  M  ,  where  the  exhaust- 
lap  line  cuts  the  ellipse.  The  crank-end  events  are  determined 
from  the  points  K,  L,  N,  and  P.  The  vertical  distance  from  the 
top  point  of  the  ellipse  to  the  head-end  steam-lap  line  is  the  head- 
end maximum  port-opening,  and  the  head-end  lead  is  given  by 
the  distance  from  the  extreme  left  point  of  the  ellipse  to  the 
head-end  steam-lap  line.  The  crank-end  maximum  port-opening 
and  lead  are  found  in  a  similar  manner. 

In  actual  use,  the  ellipse  is  rather  burdensome  because  it  takes 
considerable  time  to  construct  it.  It  is  evident  also  that  the 
crank  position  at  admission  is  not  easily  determined  with  accuracy. 
The  ellipse  is  little  used  except  in  locomotive  work. 

114.  The  Bilgram  Diagram.  —  On  the  left  of  Fig.  76,  the 
crank  and  the  eccentric  are  shown  by  C  and  E,  respectively. 
The  displacement  of  the  valve  from  mid-position  is  y.  The  dis- 
tance y  is  laid  off  perpendicular  to  the  crank,  and  a  line  is  drawn 
parallel  to  the  crank  'at  a  distance  y  from  it. 


118  ENGINES   AND    BOILERS 

At  the  right  of  Fig.  76,  this  has  been  done  for  twelve  crank 
positions.  It  is  seen  that  these  lines  all  pass  through  two  points 
P  and  P',  and  that  a  line  drawn  from  P  to  P'  passes  through  the 
center  of  the  crank  circle  and  makes  an  angle  a  with  the  hori- 
zontal. Hence  the  perpendicular  distance  from  the  points  P  or 
P'  to  the  crank  at  any  position  is  the  distance  that  the  valve 
is  from  mid-position.  The  points  P  and  P'  are  called  the  con- 
struction points  in  the  Bilgram  diagram. 

Figure  77  shows  the  application  of  the  Bilgram  diagram. 
About  the  point  P  draw  two  circles,  one  whose  radius  is  equal 
to  the  head-end  steam  lap,  and  the  other  with  a  radius  equal 


FIG.  76 

to  the  head-end  exhaust  lap.  Draw  the  crank-end  lap  circles 
with  center  at  P'.  The  crank  positions  tangent  to  these  lap 
circles  give  the  positions  at  the  different  events.  Head-end  ad- 
mission occurs  when  the  valve  is  at  a  distance  from  its  mid- 
position  equal  to  the  head-end  steam  lap,  and  when  the  valve 
is  going  away  from  its  mid-position.  The  head-end  admission 
position  shown  in  Fig.  77  fulfills  these  conditions.  At  that 
time  the  crank  is  at  a  distance  from  the  point  P  equal  to  the 
head-end  steam  lap,  and  further  motion  moves  it  farther  from  P. 
The  crank  position  at  head-end  cut-off  is  tangent  to  the  head-end 
steam-lap  circle  on  the  other  side,  i.e.,  the  valve  is  then  at  a  dis- 
tance equal  to  the  steam  lap  from  mid-position  and  further  mo- 
tion brings  the  valve  nearer  mid-position. 

The  crank  positions  for  the  other  events  are  shown  in  Fig.  77. 
Reasoning  similar  to  the  preceding  will  show  them  to  be  correct. 
It  is  customary  to  draw  the  head-end  lap  circles  about  P,  and  the 


VALVES 


119 


crank-end  lap  circles  about  Pf,  although  there  is  no  inherent 
reason  for  so  doing. 

Half  of  the  valve-travel  minus  the  steam  lap  equals  the  maxi- 
mum port  opening  if  the  valve  has  no  over-travel,  i.e.  if  it  does 
not  move  beyond  the  far  edge  of  the  port.  Therefore  the  maxi- 
mum port-opening  is  as  shown  in  the  figure;  It  is  remembered 
that  the  port  is  open  a  distance  equal  to  the  lead  when  the 
crank  is  on  dead  center.  In  other  words  the  valve  is  then  at  a 
distance  equal  to  the  steam  lap  plus  the  lead  from  its  mid-position. 
Therefore  the  perpendicular  distance  of  P  from  the  horizontal 


c,e.  t  a 


FIG.  77 

axis  is  equal  to  the  steam  lap  plus  the  lead.     The  distance  of 
the  steam-lap  circle  from  the  axis  is  then  sthe  lead. 

115.  The  Zeuner  Diagram.  —  On  the  left  of  Fig.  78,  the  valve 
displacement  y  is  laid  off  radially  on  the  crank  from  the  center 
outward.  This  gives  a  point  G  on  the  crank.  This  has  been  done 
for  twelve  crank  positions  on  the  right  of  Fig.  78  and  the  points 
connected  by  a  smooth  curve.  The  points  fall  on  the  circum- 
ferences of  two  equal  circles,  the  diameter  of  each  of  which  is 
one-half  the  valve-travel.  The  line  which  forms  the  diameters 
of  these  two  circles  makes  an  angle  a  with  the  vertical.  The 


120 


ENGINES   AND   BOILERS 


circle  whose  center  is  at  A  shows  the  movement  of  the  valve  to 
the  right  of  its  mid-position,  and  is  called  the  right  valve-circle. 
The  other  circle  is  called  the  left  valve  circle;  it  shows  the  move- 
ment of  the  valve  to  the  left  of  the  mid-position.  When  the 
crank  is  drawn  in  any  position,  the  displacement  of  the  valve 
is  given  by  the  distance  from  the  center  of  the  crank  circle  to 
the  intersection  of  the  crank  with  the  valve  circle. 

The  application  of  this  diagram  is  shown  in  Fig.  79.  With 
the  crank  on  head-end  dead  center,  the  eccentric  is  at  E,  at  an 
angle  a  to  the  right  of  the  vertical.  The  diameters  of  valve 
circles,  P—'P',  are  at  an  angle  a  on  the  other  side  of  the  vertical. 


FIG.  78 

The  extremity  P  of  the  diameter  of  the  valve  circle  is  called  the 
construction  point  in  the  Zeuner  diagram.  The  lap  circles  are 
drawn  as  shown.  The  head-end  steam-lap  circle  intersects  the 
right  valve-circle  at  the  point  T.  The  crank  position  through 
T  is  then  head-end  admission,  because  with  the  crank  in  this 
position  the  valve  is  at  a  distance  equal  to  the  steam  lap  to  the 
right  of  mid-position.  The  crank  position  at  cut-off  is  drawn 
through  the  point  K,  where  the  steam-lap  circle  intersects  the 
right  valve-circle.  Head-end  release  and  compression  occur  when 
the  head-end  exhaust-lap  circle  intersects  the  left  valve-circle. 

It  may  be  proved  by  means  of  the  similar  triangles  OWE  and 
PRO  or  by  actual  construction  that  a  line  drawn  from  A  to  B  is 
tangent  to  the  steam-lap  circle  at  W .  This  line  is  perpendicular 
to  the  diameter  of  the  valve  circles.  In  like  manner  a  line  drawn 
from  I  to  F  is  tangent  to  the  head-end  exhaust-lap  circle,  and  is 
perpendicular  to  the  diameter  of  the  valve  circles.  The  same 
thing  is  true  of  the  lines  HG  and  JD  for  the  crank  end.  It  is 


VALVES 


121 


often  better  to  draw  the  lines  A  B,  IF,  HG,  and  JD  than  to 
determine  the  crank  positions  at  the  events  by  the  intersections 
of  the  valve  circle  and  the  lap  circle. 

ON  is  equal  to  the  steam  lap  plus  the  lead,  because  the  valve 
circle  cuts  the  crank  on  dead  center  at  N.  But  ONP  is  a  right 
triangle,  since  it  is  inscribed  in  a  semicircle.  If  QS  is  drawn 
parallel  to  AW,  the  triangle  OSQ  is  a  right  triangle,  and  it  is 
similar  and  equal  to  the  triangle,  ONP.  Therefore  OS  is  equal 


Right  w/re  ctrck 


be.  c.o. 

fivnfr  circle 
Eccentric  c/n,/e 


4t.  comp 
t».rel. 


ce.  co.- 


c/rc/e 


FIG.  79 


to  the  steam  lap  plus  the  lead,  and  WS  is  equal  to  the  lead.  If 
we  then  draw  a  circle  about  Q  as  a  center,  tangent  to  the  line 
AW,  its  radius  is  the  lead.  This  is  called  the  head-end  lead 
circle.  The  crank-end  lead  circle  is  drawn  about  X  tangent  to 
HG.  The  head-end  maximum  port  opening  equals  PW,  which  is 
one-half  the  valve-travel  minus  the  steam-lap.  The  line  PK  is 
perpendicular  to  the  crank  at  cut-off,  because  PKO  is  a  right 
angle  since  it  is  inscribed  in  a  semicircle. 

The  application  of  these  valve  diagrams  to  practical  problems 
will  show  their  value.  Space  will  not  be  taken  here  to  give  the 
solutions  of  the  various  common  problems  in  which  these  dia- 
grams are  used.  The  Bilgram  and  Zeuner  diagrams  are  both 


122 


ENGINES   AND    BOILERS 


adapted  to  problems  of  valve  setting,  but  the  Bilgram  diagram  is 
the  more  convenient  for  use  in  designing  valve  gears. 

116.  Types  of  Slide-valves.  —  The  simple  D   slide-valve  has 
been  discussed  and  its  action  explained.     This  type  of  valve  is 
much  used,  but  it  has  certain  defects  which  have  been  overcome 
in  other  types.    One  of  the  defects  of  the  simple  D  valve  is  the 
large  force  necessary  to  move  it  when  high  steam  pressure  is  used. 
The  steam  pressure  on  the  back  of  the  valve  presses  it  against 
the  seat.     This  pressure  times  the  coefficient  of  friction  between 
the  valve  and  the  seat  is  the  force  that  must  be  exerted  to  move 
the  valve.     The  work  done  in  operating  the  valve  is  the  force 
times  the  distance  the  valve  is  moved.    If  either  the  force  or  the 
distance  is  decreased,  the  work  necessary  to  operate  the  valve  will 
be  lessened. 

117.  Valve  with  Pressure  Plate.  —  The  pressure  on  the  back 
of  the  valve  may  be  removed  by  putting  a  pressure  plate  above 
it,  somewhat  in  the  manner  shown  in  Fig.  80.     A  steam-tight  fit 

between  the  valve  and  the 
plate  is  made  by  strips  set  into 
s|ots  jn  the  vaive.  These  are 
pressed  up  against  the  plate 
by  springs  from  beneath,  and 
they  act  in  much  the  same  way 
as  rings  on  a  piston.  If  any 
steam  leaks  by  the  strips,  it 
may  escape  to  the  exhaust 
through  a  vent  in  the  valve. 
This  scheme  enables  us  to  remove  as  much  of  the  pressure  from 
the  back  of  the  valve  as  we  desire,  but  some  pressure  downward 
is  desirable  in  order  to  keep  the  valve  firmly  seated.  Many  flat 
slide-valves  have  pressure  plates.  Aside  from  removing  the  pres- 
sure, the  plate  does  not  affect  the  valve  in  any  way. 

118.  The  Piston  Valve.  —  Instead  of  a  flat    valve  such  as  we 
have  considered,  a  piston  valve  is  used  extensively.     Figure  81 
shows  a  form  of  this  type  where  the  valve  is  cylindrical  and  slides 
in  a  cylindrical  chamber.     It  is  readily  seen  that  it  is  perfectly 
balanced,  since  the  steam  causes  no  thrust  either   endwise  or 
on  the  seat.     Piston  valves  are  very  easy  to  operate,  but  are 
liable  to  leak  steam  when  they  become  worn.     Many  of  them 


FIG.  80 


VALVES 


123 


have  rings  similar  to  piston  rings  to  keep  this  leakage  of  steam 
to  a  minimum.  If  rings  are  used  it  is  necessary  to  bridge  the 
ports.  A  broken  ring  is  liable  to  cause  severe  damage  and  care 
must  be  exercised  to  keep  them  in  good  condition. 

In  Fig.  81,  the  steam  is  led  to  the  inside  of  the  valve,  so  that 
the  steam  lap  is  on  the  other  side  of  the  port  from  the  valves 
previously  considered.  A  valve  so  constructed  is  said  to  be  an 


FIG.  81 

indirect  valve,  or  to  have  inside  admission.  Most  piston  valves 
have  inside  admission,  but  this  is  not  a  necessity.  It  is  easier 
to  keep  the  stuffing-box  tight  against  exhaust  steam  than  against 
high-pressure  steam.  With  the  inside  admission  arrangement, 
also,  the  live  steam  has  less  surface  exposed  to  radiation.  With 
an  inside  admission  valve,  the  eccentric  follows  the  crank  by  an 
angle  of  90°  —  a.  The  valve  diagrams  studied  will  have  the  same 
form  as  before,  but  what  was  right-hand  is  now  left-hand,  i.e. 
in  the  Zeuner  diagram,  for  instance,  what  was  formerly  the  right 
valve-circle  is  now  the  left,  but  otherwise  there  is  no  change  in 
the  diagram  and  no  difference 
is  made  in  the  solution  of  a 
problem. 

119.  Double-ported  Valves. 
—  As  has  just  been  mentioned, 
the  work  required  to  move  the 
valve  is  the  product  of  the 
force  required  to  move  it  and 
the  distance  it  is  moved.  The 

distance  may  be  cut  in  half  by  making  the  valve  double-ported. 
Figure  82  shows  a  so-called  double-ported  valve.  It  is  seen  that 
for  a  certain  valve  movement  twice  the  area  for  the  passage  of 
steam  is  given  in  this  type  compared  with  a  simple  slide-valve. 
There  are  many  different  forms  of  double-ported  valves,  but  they 
are  much  the  same  in  principle  as  that  shown  in  Fig.  82. 


FIG.  82 


124  ENGINES   AND   BOILERS 

120.  The  Gridiron  Valve.  —  If  the  idea  of  a  double-ported 
valve  is  carried  a  step  farther,  we  may  get  a  very  large  aggregate 
opening  for  the  passage  of  steam  with  but  a  small  movement 
of  the  valve.  Figure  83  shows  a  gridiron  valve.  In  many  valves 
of  this  type  there  are  a  large  number  of  openings,  whereas  Fig.  83 
shows  only  three.  A  valve  of  this  character  can  have  no  exhaust 
functions,  and  separate  exhaust  valves  must  be  provided.  If  one 
exhaust  valve  takes  care  of  both  ends  of  the  cylinder,  the  engine 
is  called  a  two-valve  engine;  if  there  is  a  separate  exhaust  valve 


FIG.  83 

for  each  end,  it  is  called  a  three-valve  engine.  When  gridiron 
valves  are  used,  it  is  more  common  to  have  a  steam  valve  and 
an  exhaust  valve  for  each  end  of  the  cylinder.  The  engine 
then  has  four  valves. 

If  a  governor  is  so  constructed  that  it  regulates  the  amount 
of  steam  admitted  to  the  cylinder  by  changing  cut-off,  it  is  evi- 
dent from  the  valve  diagrams  that  the  events  of  release,  com- 
pression, and  admission  are  changed  when  cut-off  is  changed  if 
a  single  valve  is  used.  Under  some  conditions  this  is  a  serious 
defect,  and  it  makes  the  use  of  separate  steam  and  exhaust  valves 
desirable. 

121.  The  Riding  Cut-off  Valve.  —  To  utilize  the  expansive 
force  of  the  steam  in  an  engine,  it  is  necessary  to  have  an  early 
cut-off.  With  a  single  valve,  early  cut-off  will  necessitate  either 
early  release  or  early  admission.  With  an  early  cut-off  and  with 
release  near  the  end  of  the  stroke,  compression  is  bound  to  occur 
too  soon  for  satisfactory  operation  under  non-condensing  con- 
ditions. The  student  only  needs  to  draw  a  valve  diagram  to  con- 
vince himself  of  this  fact.  To  use  an  early  cut-off,  and  to  have 
at  the  same  time  reasonable  percents  of  release  and  compression, 


VALVES 


125 


a  riding  cut-off  valve  is  often  used.  There  are  several  forms  of 
riding  cut-off  valves,  but  we  shall  describe  only  one  of  them. 

Figure  84  shows  the  Myers  riding  cut-off  valve.  A  main  valve 
slides  on  the  seat  in  the  same  manner  as  an  ordinary  D  valve. 
The  steam  lap  is  made  small  so  that  the  proper  relation  exists 
between  the  events  of  admission,  release,  and  compression.  If 
the  main  valve  were  acting  alone,  cut-off  would  occur  very  late. 
To  give  an  early  cut-off,  a  rider  valve  which  controls  only  the 
event  of  cut-off  is  placed  on  the  back  of  the  main  valve.  The 
working  edge  of  the  rider  valve  effects  cut-off  when  it  matches 
with  the  edges  of  the  main  valve  at  B  and  at  D. 

In  Fig.  84  both  valves  are  shown  in  their  mid-position.  This 
would  not  occur  normallv  unless  one  of  the  valves  were  discon- 


FIG.  84 

nected  from  its  eccentric,  but  the  figure  is  drawn  in  this  manner 
to  give  a  clearer  idea  of  the  laps.  Each  valve  is  driven  by  its 
own  eccentric.  The  rider  valve  is  made  in  two  parts  and  the 
relative  position  of  the  two  parts  may  be  changed  by  revolving 
the  valve  stem.  One  part  of  the  rider  valve  is  secured  to  the 
valve  stem  by  a  right-hand  thread  and  the  other  part  by  a  left- 
hand  thread.  The  hand  wheel  to  the  left  is  arranged  so  that  by 
turning  it  the  valve  stem  is  rotated  and  the  parts  of  the  valve 
are  brought  nearer  together  or  moved  farther  apart,  thereby 
effecting  a  change  in  cut-off.  With  the  parts  farther  apart,  cut- 
off occurs  earlier.  When  the  valves  are  both  in  mid-position, 
as  shown  in  Fig.  84,  it  is  easily  seen  that  the  rider  valve  has 
negative  steam  lap  or  steam  clearance. 

To  determine  the  crank  position  at  cut-off  from  the  valve  dia- 
grams, it  is  necessary  to  consider  the  relative  motion  of  the  two 
valves.  Figure  85  is  the  Zeuner  analysis  for  the  rider  valve. 
The  crank  circle  and  the  two  eccentric  circles  are  shown  by  the 
light  lines.  The  point  PI  is  the  extremity  of  the  diameter  of  the 


126 


ENGINES   AND    BOILERS 


right  valve-circle  for  the  main  valve,  and  a\  is  the  angle  of  ad- 
vance for  the  main  eccentric.  The  point  P%  is  the  extremity  of 
the  diameter  of  the  right  valve-circle,  and  a2  is  the  angle  of  advance 
for  the  rider-valve  eccentric.  If  a  number  of  crank  positions  be 
chosen  and  the  displacement  of  the  rider  valve  relative  to  the 
main  valve  be  laid  off  on  the  crank  from  the  center  radially 
outward,  a  number  of  points  will  be  established.  Connecting 


Relative 
ffiqht  l/e/ve  Ctrt/e 


C.e.  f?/cfer  fa/re 
Steam  top  (neyafiri) 


ff/JfrVafoe 
c.e  a/t-o/f. 


FIG.  85 

these  points  by  a  smooth  curve,  we  find  that  it  is  composed  of 
the  two  circles  shown  by  heavy  lines  in  Fig.  85.  The  point  PS 
is  the  extremity  of  the  diameter  of  the  right  relative  valve-circle 
and  «3  is  the  angle  which  that  diameter  makes  with  the  vertical. 

It  is  seen  that  the  diameter  of  the  right  relative  valve-circle, 
OPs,  is  equal  in  amount  and  parallel  to  the  dotted  line  PiP2. 
Knowing  this  fact,  the  solution  of  a  rider-valve  problem  is  quite 
simple,  for  it  is  not  necessary  to  plot  the  points  to  determine 
the  relative  valve-circle. 

Since  the  steam  lap  for  the  rider  valve  is  negative,  the  intersec- 
tion of  that  lap  circle  with  the  left  valve-circle  gives  the  crank 
position  at  head-end  cut-off.  The  cut-off  positions  of  the  crank 
are  shown  by  the  heavy  lines  in  Fig.  85.  The  crank  positions 
for  admission,  release,  and  compression  are  determined  from  a 
valve  diagram  for  the  main  valve  in  exactly  the  same  manner  as 
previously  explained  for  the  D  valve. 

122.  Effect  of  Rocker  Arm  on  Location  of  Eccentric.  —  In 

the  previous  discussion  of  slide-valves,  it  has  been  supposed  that 


VALVES 


127 


the  eccentric  rod  is  attached  directly  to  the  valve  rod,  and  that 
the  movement  of  the  valve  is  the  same  as  the  horizontal  move- 
ment of  the  eccentric.  Quite  often  a  rocker  arm  is  interposed 
between  the  eccentric  and  the  valve  rods,  in  which  case  it  may 
be  necessary  to  modify  our  previous  assumption. 

In  Fig.  86,  three  arrangements  of  rocker  arms  are  shown.  In 
I,  the  eccentric  rod  and  the  valve  rod  are  both  connected  to  the 
same  pin  at  B,  and  our  assumption  is  not  changed. 

In  II,  the  arm  B  AC  reverses  the  motion  of  the  eccentric.  In 
this  case,  if  a  direct  valve  is  used,  the  eccentric  must  be  placed 
on  the  shaft  at  180°  from  the  position  it  would  have  had  without 
the  rocker  arm,  i.e.  if  a  direct  valve  is  used,  the  eccentric  must 
follow  the  crank  by  an  angle  of  90°  —  a. 

If  an  indirect  valve  is  used  with  a  reversing  rocker,  the  eccen- 


Eccentr/e  /foe/ 


Eccentric  flotf 


n 
FIG.  86 


m 


trie  is  placed  90°+  a  ahead  of  the  crank.     This  modification  does 
not  effect  the  work  in  making  the  valve  analysis. 

In  case  III,  the  two  arms  of  the  rocker  AB  and  AC  are  not 
of  the  same  length.  The  travel  of  the  valve  is  the  diameter  of 
the  eccentric  circle  times  AB/AC. 

123.  Oscillating  Valves.  —  One  of  the  most  common  valves  is 
cylindrical  and  oscillates  or  rocks  on  a  cylindrical  seat.  A  spindle 
fastened  to  the  valve  extends  out  of  the  steam  chest  and  carries 
an  arm  that  is  moved  back  and  forth  by  the  eccentric.  The 
Corliss  valve  shown  in  Fig.  58  and  the  valves  of  Figs.  60  and  61 
are  of  this  type.  It  would  be  possible  to  make  an  oscillating 
valve  control  both  the  admission  and  exhaust  events,  but  this  is 
seldom  done.  Where  the  oscillating  valve  is  used,  four  valves 
usually  are  employed.  Because  of  the  distortion  of  motion  due 
to  the  valve  arm,  it  is  not  possible  to  show  by  the  valve  diagrams 


128 


ENGINES   AND    BOILERS 


the  exact  motion  of  the  valve.  However,  the  relation  of  horizontal 
motion  of  the  eccentric  still  holds,  and  so  the  valve  diagrams  are 
of  value  in  the  analysis  of  these  valves. 

124.  Poppet  Valves.  —  While  they  are  not  common  for  steam 
engines  in  this  country,  most  gasoline  engines  are  equipped  with 
poppet  valves.  This  is  a  lifting  valve  and  there  is  no  sliding  of 
the  valve  on  the  seat.  Figure  87  shows  this  type  of  valve.  These 
valves  do  not  have  to  be  lubricated  and  would  seem  to  be  well 
adapted  to  conditions  where  highly  superheated  steam  is  used, 
since  one  of  the  troubles  met  with  in  the  use  of  superheated  steam 
is  the  difficulty  of  proper  lubrication  of  the  valve.  Where  poppet 

valves  are  used  on  steam  engines, 
they  usually  are  operated  either  by  a 
cam  or  by  an  eccentric  from  a  lay 
shaft  that  is  parallel  to  the  axis  of 
the  cylinder.  The  lay  shaft  usually 
is  driven  by  mitre  gears  from  the 
main  shaft. 

125.   Reversing.  —  The  direction  of 
rotation  of  an  engine  may  be  changed 
FIG.  87  by  shifting  the  eccentric  to  the  proper 

position.      Occasionally    an    ordinary 

engine  must  be  reversed.  Unless  the  eccentric  is  keyed  to  the 
shaft,  this  is  not  usually  a  difficult  task.  With  a  direct  valve, 
the  eccentric  leads  the  crank  by  an  angle  of  90°+  a.  This  is  true 
irrespective  of  the  direction  of  rotation.  To  reverse,  then,  move 
the  eccentric  in  the  direction  the  engine  has  been  running  through 
an  angle  of  180°—  2  a.  The  same  rule  applies  with  an  indirect  or 
inside-admission  valve. 

With  certain  classes  of  engines,  such  as  are  used  for  locomotives, 
in  marine  work,  etc.,  reversing  is  a  common  occurrence.  Some 
handier  and  quicker  means  must  then  be  provided  than  that  men- 
tioned above.  The  devices  used  for  this  purpose  are  called  re- 
versing gears.  There  are  a  very  great  many  types  of  reversing 
gears  in  use,  but  space  will  not  permit  the  discussion  of  more 
than  the  most  common  types. 

126.  The  Stephenson  Link.  —  One  of  the  most  widely  used 
reversing  gears  is  the  Stephenson  link  gear,  which  is  much  used 
for  small  locomotives.  In  this  arrangement,  there  are  two  ec- 


VALVES 


129 


Gentries  placed  on  the  shaft  at  an  angle  of  180°  —  2  a  apart.  The 
forward  eccentric  controls  the  forward  motion,  and  the  back- 
ward eccentric  controls  the  reverse  motion.  In  the  diagram  of 
Fig.  88,  the  crank  is  shown  on  head-end  dead  center,  and  the 
valve  is  indirect,  so  that  the  eccentrics  will  be  at  an  angle  of  90°  —  a 


FIG.  88 

from  the  crank.  In  Fig.  88,  FE  is  the  forward  eccentric,  and 
BE  the  backward  eccentric.  The  two  eccentric  rods  connect  to 
the  eyes  of  a  link  at  H  and  I.  This  link  may  be  raised  or  low- 
ered by  the  bell-crank  BAD  and  the  link  DJ.  When  the  link 
is  down,  as  shown  in  Fig.  89,  the  forward  eccentric  entirely  con- 
trols the  motion  of  the  valve.  With  the  link  all  the  way  up,  the 
backward  eccentric  controls  the  valve.  In  Fig.  88,  the  link  is 


FIG.  89 

shown  in  mid-position  with  both  eccentrics  controlling  an  equal 
amount.  With  the  crank  on  head-end  dead  center  as  in  Fig.  88, 
the  valve  is  as  far  to  the  left  as  it  can  get,  the  port  is  open  a  dis- 
tance equal  to  the  lead,  and  the  valve  is  at  a  distance  equal  to 
the  steam  lap  plus  the  lead  from  its  mid-position.  With  the 


130 


ENGINES   AND    BOILERS 


valve  opening  only  lead  distance,  the  engine  will  not  get  enough 
steam  to  run.  At  mid-gear,  half  the  travel  of  the  valve  is  equal 
to  the  steam  lap  plus  the  lead. 

With  the  link  in  some  position  between  those  shown  in  Figs.  88 
and  89,  both  eccentrics  will  control  the  motion  of  the  valve,  but 
the  forward  one  will  predominate,  and  the  engine  will  run  for- 
ward. The  cut-off  will  now  be  earlier  than  it  would  be  if  the 
link  was  all  the  way  down  in  full  gear.  It  is  evident  that  the 
Stephenson  gear  may  be  used  to  change  the  cut-off  as  well  as  to 
reverse. 

It  is  possible  to  give  the  same  motion  to  the  valve 
with  the  link  in  intermediate  position,  by  a  simple  equivalent 
eccentric.  The  method  of  determining  this  equivalent  eccentric 
is  not  difficult  but  it  will  not  be  discussed  here. 

127.  The  Walschaert  Valve  Gear.  —  Most  large  locomotives 
in  this  country  are  equipped  with  the  Walschaert  gear,  or  some 


FIG.  90 


similar  reversing  gear.  In  this  type  of  gear,  shown  in  Fig.  90, 
the  motion  of  the  valve  is  derived  partly  from  the  cross-head, 
and  partly  from  a  return  crank  or  eccentric.  That  part  of  the 
motion  coming  from  the  cross-head  is  constant,  while  that  de- 
rived from  the  eccentric  is  varied  for  different  conditions  of  load 
and  direction  of  rotation.  Since  the  eccentric  is  placed  outside 
the  driver,  it  is  commonly  called  a  return  crank.  The  bar  CE 
is  fastened  to  the  outer  end  of  the  crank  pin  C,  and  the  eccentric 
pin  E  moves  in  the  dotted  circle  (Fig.  90),  about  0  as  a  center. 


VALVES  131 

The  angle  between  the  crank  and  the  eccentric  is  90°.  The  hori- 
zontal motion  of  the  eccentric  is  transmitted  through  the  eccen- 
tric rod  EM  to  the  lower  point  of  a  link.  The  link  is  pivoted 
at  its  center  G  to  the  frame  of  the  engine,  so  that  the  point  M 
oscillates  about  the  point  G.  In  this  link  is  fitted  a  block  which 
can  be  raised  or  lowered  in  the  link  by  the  bell  crank  DAB. 
As  shown  in  the  dotted  position,  the  block  is  at  its  lowest  posi- 
tion in  the  link,  and  the  engine  is  running  forward,  taking  steam 
during  the  largest  possible  part  of  stroke.  In  the  full  line  posi- 
tion, the  block  is  at  the  center  of  the  link,  and  the  engine  will 
not  get  enough  steam  to  drive  it.  With  the  block  in  mid-position 
in  the  link,  the  eccentric  will  give  no  motion  whatever  to  the 
valve.  Under  this  condition  the  motion  of  the  valve  comes 
entirely  from  the  cross-head. 

The  lever  HIJ  is  called  the  combination  lever,  because  it  com- 
bines the  motion  from  the  eccentric  with  the  motion  from  the 
cross-head.  The  ratio  I H/JH  is  fixed  by  the  condition  that 

I H  __  2(steam  lap  plus  lead) 
J  H  length  of  stroke 

With  the  block  in  the  center  of  the  link  at  G,  H  has  no  horizontal 
motion,  but  the  horizontal  motion  of  /  is  equal  to 

length  of  stroke  X  I H     n  ,  ,         ,        ,      ,      ,. 

• T-PT -  =  2  (steam  lap  plus  lead). 

J  ti 

As  the  motion  of  the  valve  at  mid-gear  is  2 (steam  lap  plus  lead), 
it  is  seen  that  the  valve  will  open  only  a  distance  equal  to  the 
lead  on  each  end,  and  the  engine  will  not  get  enough  steam  to 
run.  When  the  block  is  dropped  to  the  dotted  position,  the  point 
H  does  have  a  horizontal  motion  which  comes  from  the  eccen- 
tric, and  with  the  engine  running  in  the  direction  shown,  the 
horizontal  motion  of  H  will  add  to  the  port  opening.  To  reverse, 
the  block  is  raised  above  the  center  of  the  link. 

By  changing  the  position  of  the  block  in  the  link  we  may  get 
not  only  a  reversal  of  direction  of  rotation  but  also  a  change  in 
cut-off.  As  in  the  Stephenson  gear,  it  is  possible  to  find  an  equiv- 
alent eccentric  which  would  give  the  motion  actually  obtained 
from  the  mechanism. 


132  ENGINES   AND   BOILERS 

128.  The  Joy  Valve  Gear.  —  The  Joy  gear  is  a  so-called  radial 
gear.  Unlike  those  previously  described,  it  has  no  eccentric. 
Figure  91  shows  diagrammatically  the  principle  of  its  operation. 
A  point  such  as  D  on  the  connecting  rod  FC  will  move  in  the 
path  of  an  ellipse,  which  is  shown  dotted.  A  bar  BED  is  pinned 
to  the  connecting  rod  at  D.  The  other  end  of  the  bar  is  con- 
nected to  the  frame  of  the  engine  by  the  link  AB,  A  being  a  point 
on  the  engine  frame.  It  is  evident  that  a  point  E  on  this 
bar  will  have  a  combination  of  the  elliptic  motion  of  D 
and  the  nearly  vertical  motion  of  B.  The  path  of  E  is  shown 
dotted.  A  bar  EGH  is  connected  to  BED  by  the  pin  E.  The 
point  G  is  in  a  block  which  is  at  liberty  to  slide  along  a  curved 


FIG.  91 

link.  At  any  particular  cut-off  this  link  is  held  stationary,  and 
the  block  G  slides  up  and  down  in  it.  It  is  thus  seen  that  the 
point  H  gets  a  combination  of  the  motions  of  the  point  E  and 
of  the  point  G.  The  horizontal  component  of  the  motion  of  H 
is  transmitted  through  the  link  I H  to  the  valve.  By  rotating 
the  link  about  its  center  J,  a  reversal  in  the  direction  of  rotation 
of  the  shaft  may  be  obtained.  As  in  the  other  gears,  cut-off 
may  be  varied  as  well  as  the  engine  reversed. 

Under  conditions  of  light  load  an  early  cut-off  may  be  used 
and  sufficient  power  obtained  if  the  steam  is  not  throttled  between 
the  boiler  and  the  engine.  It  is  the  practice  on  locomotive  engines 
to  vary  the  cut-off  to  suit  the  load  under  normal  running  condi- 
tions. The  power  generated  by  the  engine  might  also  be  regu- 
lated by  throttling  the  steam,  but  it  has  been  found  that  a  higher 
efficiency  is  obtained  at  light  loads  by  using  an  early  cut-off 


VALVES  133 

than  by  using  a  late  cut-off  and  a  low  steam  pressure,  especially 
where  valve  gear  is  used  which  increases  compression  when  cut- 
off is  shortened,  as  in  the  Stephenson  gear.  In  the  reversing 
gears  commonly  used,  an  early  cut-off  is  accompanied  by  an  early 
compression.  The  increased  efficiency  at  light  loads  is  due,  how- 
ever, more  to  the  early  cut-off  than  to  the  high  compression, 
although  the  high  compression  aids  by  heating  the  clearance 
space,  piston,  and  cylinder  head,  thereby  keeping  initial  conden- 
sation within  more  reasonable  limits. 

129.  Setting  the  Slide-valve.  —  On  a  small  engine  that  can 
be  turned  over  easily  by  hand,  the  setting  of  a  slide-valve  is  a 
simple  matter.  With  proper  valve  setting  an  engine  should  run 
smoothly,  should  be  easy  to  start,  and  each  end  of  the  cylinder 
should  furnish  about  half  the  power.  If  an  indicator  is  at  hand, 
it  should  be  employed  in  the  setting.  If  no  indicator  is  available, 
the  valve  may  be  set  by  linear  measurement.  In  the  setting  of 
a  slide-valve  there  are  but  two  things  to  do,  shift  the  eccentric  on 
the  shaft,  and  lengthen  or  shorten  the  valve  stem  or  rod. 

VALVE  SETTING  BY  INDICATOR.  In  order  to  get  approximately 
the  same  amount  of  work  from  each  end  of  the  cylinder,  cut-off 
for  the  two  ends  should  be  about  the  same.  A  rough  adjust- 
ment may  be  made  easily  by  placing  the  eccentric  somewhere 
near  its  proper  position,  and  adjusting  the  position  of  the  valve 
on  the  rod  so  that  the  engine  may  be  started.  After  the  engine 
is  started,  take  cards  and  then  adjust  the  length  of  the  valve 
stem  until  the  cut-off  is  the  same  percent  on  both  ends.  Next, 
shift  the  eccentric  until  the  desired  percentage  of  cut-off  is  attained. 
Shifting  the  eccentric  ahead  makes  cut-off  come  earlier.  By  the 
use  of  the  indicator  it  is  easy  to  get  the  exact  setting  desired. 

Knowing  the  valve-travel  and  the  dimensions  of  the  valve,  we 
may  compute  by  means  of  the  Zeuner  valve  diagram  the  exact 
amount  the  valve  stem  must  be  lengthened  or  shortened  and  the 
angle  the  eccentric  must  be  shifted  to  give  a  desired  setting. 
In  the  upper  part  of  Fig.  92  are  shown  two  cards  taken  with  a 
valve  as  now  set.  It  is  desired  to  change  the  setting  so  that 
cut-off  will  be  50  per  cent  for  each  end.  The  cards  are  shown 
in  length  equal  to  the  valve-travel,  but  they  need  not  have  been, 
as  the  percentages  of  stroke  could  have  been  scaled  and  the 
corresponding  crank  positions  found. 


134 


ENGINES   AND    BOILERS 


From  the  cards,  locate  on  the  crank  circle  (assumed  for  con- 
venience with  its  diameter  the  same  as  that  of  the  valve-travel 
circle),  the  crank  positions  at  the  different  events.  Draw  lines 
between  the  crank  positions  at  admission  and  cut-off,  and  from 
release  to  compression,  for  each  end.  The  distance  between  the 
admission-cut-off  lines  should  be  the  sum  of  the  steam  laps  as 
measured  on  the  valve  itself.  A  radial  line  drawn  perpendicular 
to  the  line  joining  admission  and  cut-off  establishes  the  angle  of 
advance  a\.  The  laps  may  be  measured  from  the  diagram  as 
shown.  Now  construct  a  Zeuner  diagram  (Fig.  93)  for  the  desired 


c.e.ad/n 


FIG.  92 


FIG.  93 


cut-off  (50  per  cent  in  the  diagram)  keeping  the  sum  of  the  steam  laps 
the  same  as  before.  This  amount,  and  also  the  sum  of  the  steam 
lap  and  the  exhaust  lap  for  each  end,  will  be  the  same  no  matter 
what  adjustment  is  made.  The  angle  of  advance  for  the  new 
condition  is  0:2.  Then  <*2— a.\  is  the  angle  that  the  eccentric  must 
be  shifted  forward.  The  difference  X  between  the  new  and  the 
old  head-end  steam  laps  is  the  distance  the  valve  must  be  moved 
on  the  rod.  Since  the  new  head-end  steam  lap  is  larger  than  the 
old,  the  valve  rod  must  be  lengthened  if  the  valve  is  direct,  and 
shortened  if  the  valve  is  indirect.  The  upper  part  of  Fig.  93  shows 
the  cards  that  may  be  expected  after  the  setting  has  been  changed. 


VALVES  135 

VALVE  SETTING  BY  MEASUREMENT.  If  an  indicator  is  not  avail- 
able, the  valve  may  be  set  by  linear  measurement.  The  steam  chest 
cover  must  be  removed  so  that  the  measurements  may  be  taken. 
It  is  customary  to  set  either  for  equal  cut-offs  or  for  equal  leads. 
We  cannot  have  equal  leads  and  equal  cut-offs  at  the  same  set- 
ting because  of  the  angularity  of  the  connecting  rod.*  In  either 
case  adjust  the  valve  on  the  stem  so  that  the  valve  travels  about 
as  far  beyond  the  head-end  port  as  beyond  the  crank-end  port. 

SETTING  FOR  EQUAL  CUT-OFF.  By  means  of  marks  on  the  guides 
and  on  the  cross-head  the  stroke  may  be  determined,  and  that 
proportion  from  each  end  at  which  cut-off  is  to  take  place  may 
be  laid  off  on  the  guides.  The  procedure  is  as  follows: 

(1)  Place  the  cross-head  at  the  position  for  head-end  cut-off. 

(2)  Loosen  the  eccentric  on  the  shaft  and  turn  it  on  the  shaft  in 
the  direction  the  engine  is  to  run  until  the  valve  is  just  on  the  point 
of  cutting  off.     Fasten  the  eccentric  to  the  shaft  in  this  position. 

(3)  Turn  the  engine  to  the  position  for  cut-off  at  the  crank-end 
and  measure  the  distance  from  the  valve  to  its  correct  position 
to  give  cut-off  on  this  end.     Divide  the  error  by  two  and  take 
up  half  the  error  by  shifting  the  valve  on  the  stem  and  the  other 
half  by  turning  the  eccentric  on  the  shaft. 

(4)  Turn  the  engine  back  to  the  position  for  head-end  cut-off 
and  check  the  setting.     If  there  is  an  error  left,  repeat  as  ex- 
plained above. 

Be  sure  the  eccentric  is  fastened  firmly  to  the  shaft  and  the 
valve  to  the  stem.     Replace  the  cover  of  the  steam  chest. 
SETTING  FOR  EQUAL  LEAD. 

(1)  Place  the  engine  accurately  on  head-end  dead  center. 

(2)  Loosen  the  eccentric  on  the  shaft  and  turn  it  in  the  direc- 
tion the  engine  is  to  run  until  the  port  is  open  a  distance  equal  to 
the  desired  lead.    Be  sure  the  valve  will  open  the  port  if  the  eccen- 
tric is  turned  ahead  more.     Fasten  the  eccentric  to  the  shaft. 

(3)  Turn  the  engine  to  crank-end  dead  center  and  measure  the 
error.     Divide  the  error  by  two  and  take  up  half  by  turning  the 
eccentric  on  the  shaft  and  the  other  half  by  adjusting  the  length 
of  the  valve  stem. 

(4)  Turn  the  engine  back  to  head-end  dead  center  and  check. 
If  there  is  an  error,  correct  it  by  repeating  as  previously  explained. 

*  This  expression  means  the  deviation  from  parallelism  with  the  axis  of  the  cylinder,  of  a 
connecting  rod  of  finite  practical  length,  except  when  the  crank  is  at  one  of  the  dead  centers. 


136 


ENGINES   AND    BOILERS 


The  reason  that  half  the  error  is  taken  up  by  moving  the  valve 
on  the  stem  and  half  by  turning  the  eccentric  on  the  shaft  may 
be  explained  as  follows.  Suppose  the  valve  has  been  set  to  give  the 
correct  cut-off  on  the  crank-end,  but  that  when  turned  over  to 
head-end  position  for  cut-off  there  is  an  error  as  shown  in  Fig.  94. 

The  eccentric  is  now  at  E\. 
By  turning  the  eccentric  from 
Ei  back  to  E2)  this  error 
would  be  adjusted,  but  nothing 
would  have  been  gained  as  will 
be  seen  when  the  engine  is 
turned  back  to  the  position  for 
crank-end  cut-off.  The  same 
error  now  exists  on  crank-end  as  shown  in  Fig.  95.  That  is,  the 
eccentric  is  at  E*  and  should  be  at  E%.  If  we  try  to  take  up  the 
error  by  lengthening  the  valve  stem  (Fig.  96),  nothing  is  gained 
because  the  valve  will  be  moved  to  the  left  a  distance  equal  to  the 
error,  and  when  it  is  turned  back  to  the  position  for  crank-end  cut- 
off, the  valve  will  be  open  a  distance  equal  to  the  error.  If  now 
we  divide  the  error  by  two,  and  move  the  edge  of  the  valve  to  A, 


FIG.  94 


FIG.  95 


FIG.  96 


in  Fig.  94,  by  turning  the  eccentric  back  from  EI  to  D,  we  will 
find  that  the  other  edge  of  the  valve  will  be  at  B  in  Fig.  95,  when 
turned  over  to  the  crank-end  cut-off  position.  If  the  other  half 
of  the  error  be  taken  up  in  Fig.  94  by  lengthening  the  valve  stem 
i.e.  by  moving  the  valve  to  the  left  on  its  stem,  the  valve  will  be 
in  the  correct  position  for  head-end  cut-off.  Moreover,  moving 
the  valve  to  the  left  moves  the  crank-end  edge  (Fig.  95),  from  B  to 
the  edge  of  the  port,  where  it  should  be  for  crank-end  cut-off. 
Therefore,  by  taking  up  half  the  error  in  each  way  at  the  posi- 
tion of  head-end  cut-off,  we  have  made  no  net  change  in  the 
position  of  the  valve  for  the  crank-end  cut-off  position. 


CHAPTER  IX 
GOVERNORS 

130.  General.  —  The  function  of  the  governor  is  to  keep  the 
engine  running  at  nearly  the  same  speed  at  all  loads.  It  does 
this  by  controlling  the  amount  of  steam  admitted  to  the  cylinder. 
It  is  impracticable  for  the  governor  to  keep  the  speed  exactly 
constant  at  all  loads.  This  may  be  seen  when  we  understand 
how  a  governor  must  work.  Under  conditions  of  changing  load, 
the  governor  must  change  the  amount  of  steam  admitted  to  do 
the  work.  At  any  instant  we  have  a  certain  load  on  the  engine, 
the  governor  is  admitting  enough  steam  to  carry  that  load.  If 
an  extra  load  is  thrown  on,  the  engine  will  momentarily  slow 
down.  This  slower  speed  affects  the  position  of  the  parts  of  the 
governor,  and  this  in  turn  allows  more  steam  to  enter  to  do  the 
extra  work  required.  This  change  cannot  be  affected  instantane- 
ously, although  some  governors  respond  in  a  very  short  time. 
Governors  that  are  quick  in  making  the  change  are  said  to  be 
sensitive,  and  those  that  are  slow,  sluggish. 

Under  full-load  conditions,  the  engine  usually  runs  slower  than 
that  at  no  load,  by  an  amount  that  depends  upon  the  construction 
of  the  governor.  Governors  are  made  that  give  an  engine  speed 
which  is  nearly  as  great  at  full  load  as  at  no  load.  These  are  said 
to  give  doss  regulation.  It  is  possible  to  design  and  construct 
a  governor  that  gives  the  same  average  engine  speed  at  all  loads, 
in  which  case  the  governor  is  said  to  be  isochronous.  Such  a 
governor  would  tend  to  hunt,  i.e.  there  would  be  a  constant 
fluctuation  in  speed  as  the  governor  attempted  to  regulate  the 
steam  supply  to  balance  the  load  conditions. 

Practical  considerations  limit  the  nearness  to  which  isochronism 
may  be  approached.  If  the  no-load  speed  is  greater  than  the 
full-load  speed  the  governor  is  said  to  be  stable.  If  the  governor 
is  isochronous  or  gives  a  full-load  speed  greater  than  no-load 
speed,  it  is  unstable.  An  unstable  governor  is  clearly  undesir- 
able. Various  schemes  are  used  to  express  the  relation  of  speeds 
at  various  loads.  A  common  way  is  to  express  the  variation  of 
speed  from  no  load  to  full  load  and  from  no  load  or  full  load  to 
normal  load  in  percent  of  the  speed  at  normal  load.  We  may 

137 


138  ENGINES   AND    BOILERS 

then  say,  the  percent  of  variation  in  speed  from  no  load  to  full 
load  is  equal  to  100(ni  —  n^/n,  where  HI  and  n%  are  the  speed  at  no 
load  and  the  speed  at  full  load,  respectively,  and  n  is  the  speed 
at  normal  load. 

131.  Classification  of  Governors.  —  Governors  may  be  classi- 
fied according  to  the  following  characteristics. 

(a)  As  to  the  manner  of  regulating  the  steam  supply:    Under 
this  head  we  have  (1)  throttling  governors,  which  regulate  the 
amount  of  steam  admitted  to  the  cylinder  by  controlling  a  throttle 
valve,  and  (2)  cut-off  governors,  which  control  the  steam  supplied 
by  changing  the  point  of  cut-off. 

(b)  As  to  the  predominant  controlling  force  in  the  mechanism: 
We  speak  of  centrifugal  governors,  inertia  governors  (although 
inertia  is  not  a  force),  and  resistance  governors.     All  mass  has 
inertia.     If  the  mass  of  the  moving  parts  is  small  and  the  inertia 
effect  is  not  used  in  governing,  we  call  the  governor  a  centrifugal 
governor.     If  the  inertia  is  large  and  its  effect  is  used  to  aid  in 
governing,  we  have  what  we  call  inertia  governors.    Even  in  inertia 
governors,  the  centrifugal  force  is  a  very  important  factor. 

(c)  As  to  the  force  used  to  balance  the  centrifugal  force  of  the 
rotating  parts*:   We  have  gravity-balanced  governors  and  spring- 
balanced  governors. 

(d)  As   to   the   arrangement   of  the   mechanism:    There  are 
spindle  governors  and  shaft  governors. 

132.  The  Gravity-balanced  Spindle  Governor.  —  The  diagram 
of  Fig.  97  represents  a  simple  gravity-balanced  spindle  governor. 
This  is  sometimes  called  a  conical  pendulum,  and  is  also  often 
called  the  Watt  governor,  because  James  Watt  first  used  it  on 
his  engines.     Two  flyballs,  at  the  ends  of  arms,  rotate  about  a 
vertical  spindle.     The  arms  are  pivoted  to  the  spindle  at  0.     The 
height  of  the  balls  is  caused  to  control  the  steam  supply.     In 
Fig.  97  this  is  done  by  raising  or  lowering  the  point  A  with  the  balls, 
which,  by  a  suitable  mechanism,  causes  either  the  movement  of 
a  throttle  valve  or  a  change  in  the  point  of  cut-off. 

A  definite  relation  exists  between  the  height  hi  of  the  cone  of 
revolution,  and  the  speed  of  the  spindle.  To  determine  this  rela- 
tion consider  one  of  the  balls  as  a  free  body.  At  any  certain 
speed  it  may  be  considered  to  be  in  equilibrium  under  the  action 
of  the  following  forces:  the  tension  in  the  arm  T,  the  weight  W, 


GOVERNORS 


139 


and  the  centrifugal  force  acting  radially  outward  (W/g)X(vz/R). 
Taking  moments  of  these  forces  about  the  point  0,  we  have 

— X^Xh-WXR  =  0. 
g      ti 

But  v  =  2irRn,  where  n  is  the  number  of  revolutions  per  unit  time. 
Therefore  we  may  write 


1  =WR 


gR 


or 


If  we  wish  to  express  hi  in  inches  and  the  speed  of  the  governor 
in  r.  p.  m.,  our  equation  becomes 


hi  — 


32.2X12X3600      35200 
-        .   9    9  —  —  z—j 

4?r2  n2  n2 


approximately. 


From  this  equation  it  is  seen  that  the  height  hi  of  the  cone  of  rev- 
olution does  not  depend  upon  the  length  of  the  arm.     Figure  97 


FIG.  97 

shows  the  r.  p.  m.  corresponding  to  the  height  /ii  for  two-inch 
increments  of  /i,  up  to  20  inches.  From  these  values  it  will  be 
noticed  that  for  a  certain  vertical  movement  of  the  ball  there  will 
be  less  speed  variation  with  the  ball  in  the  lower  positions.  In 
other  words,  to  get  a  reasonable  speed  variation  it  will  be  neces- 
sary to  run  the  governor  very  slowly.  At  low  speeds  the  gov- 
ernor will  not  have  much  power  unless  the  balls  are  made  exces- 
sively heavy.  This  practical  limitation  precludes  the  use  of  this 
governor  on  modern  engines. 


140 


ENGINES   AND   BOILERS 


FIG.  98 


If  the  governor  arms  are  crossed,  as  shown  in  Fig.  98,  it  will 
be  noticed  that  much  less  variation  in  speed  exists  for  the  same 
vertical  movement  of  the  balls  than  for  the  form  shown  in  Fig.  97. 
Moreover,  this  governor  is  nearly  isochronous  at  50  r.  p.  m. 

By  the  proper  selec- 
tion of  the  pivots  B 
and  C,  we  may  get 
quite  satisfactory 
speed  regulation  for  a 
certain  limited  range 
of  vertical  movement 
at  any  desired  speed. 

The  pendulum  gov- 
ernor may  be  made 
exactly  isochronous  by 
making  the  balls  swing  up  in  the  arc  of  a  parabola,  as  in  Fig.  99. 
The  subnormal  of  a  parabola  is  constant,  and  it  is  seen  in  Fig.  99 
that  hi  is  the  sub-normal  of  the  parabola  which  is  the  path  of 
the  balls  as  they  swing  upward.  The  balls  may  be  made  to  take 
the  parabolic  path  by  having  the  arms  made  flexible  and  to  unwind 
from  the  evolute  of  the  parabola  or  by  having  them  guided  by 
an  arrangement  of 
cams.  Of  course  it  is 
understood  that,  in 
practice,  the  governor 
never  would  be  made 
exactly  isochronous, 
but  it  is  seen  that  iso- 
chronism  may  be  ap- 
proached as  nearly  as 
practical  conditions 
will  permit. 

In  order  to  run  a 
spindle  governor  at 
fairly  high  speed,  and  FlG>  99 

still  have  a  reasonably 

small  speed  variation  in  the  engine,  it  is  customary  to  load  it  as 
shown  in  Fig.  100.  The  load  L  tends  to  pull  the  balls  down; 
hence  they  must  rotate  faster,  to  get  to  the  same  height  as 
before.  To  find  the  relation  between  the  height  of  the  cone  of 


GOVERNORS 


141 


revolution  and  the  speed,  consider  the  load  and  the  ball  each  as 
a  free  body.  With  the  forces  acting  on  them  as  shown,  we  can 
express  the  conditions  of  equilibrium  as  follows.  Expressing  the 
fact  that  the  sum  of  the  vertical  forces  is  zero  for  the  load  L, 


(1) 
or 


2T2  sin  j8  =  L, 

T        -±- 

2sin/3 


Considering  the  ball  as  a 
free  body,  and  taking 
moments  about  0  as  a 
center,  we  have,  since  the 
sum  of  these  moments  must 
be  zero, 


W  v2 
(2)         -p 

Q   rt 


-WXR-  !F2sin 
By  (1)  this  may  be  written  in  the  form 
(3)  ~^Xhi  ~WR~  f  X#  -  |  ctn 

or,  since  v  =  2irnR, 

W  T  T 


0. 


(4)  j  WRtfh,  -WR-^R-^ctu 

whence,  since  ctn  /3  =  R/h2, 

—    TT     n    l- 
and  finally,  solving  for  n2, 

r^+K^+A.)]    ,1  + 

(6)          »*  =  i- 

If  /Z-i  =  ^2 


o, 


=  0, 


=  0, 


L1)J 


7 


or 


fe 


/Tf  +ZA 
V     IT     / 


If  hi  be  expressed  in  inches  and  n  in  r.  p.  m., 

W  +  L\  35200 


142 


ENGINES   AND   BOILERS 


The  usual  form  of  loaded  governor  is  shown  in  Fig.  101.  This 
is  seen  to  vary  slightly  from  Fig.  100.  Taking  the  load  as  a  free 
body,  T2  =  L/(2  sin  /3),  as  in  (1)  above. 


FIG.  101 


Taking  the  upper  arm  as  a  free  body,  and  expressing  the  fact 
that  the  sum  of  the  moments  about  0  equal  to  0, 


(9)      —  Xj^XH  -WXR- 


(10)       -X^XH- 

g    R 


in  PX^R-  T2  cos  ftXhi  =  0. 
Substituting  the  value  of  T2  from  (1),  we  have 

6  2   ^b  2 

But  we  have 

(ID 

hence 

(12)        ~X^H  - 

Since  we  have  also  v  =  2irRn, 


hi  =  H  T       and 
o 


_w__  _ 

2  b       2/i 


whence 

' 


. 


GOVERNORS  143 

Taking  as  our  units  inches  and  r.  p.  m.,  this  becomes 

(15)  n2  = 


H 

In  the  solution  of  a  problem  for  this  type  of  governor  it  is  best 
to  make  a  drawing  and  to  scale  from  it  the  values  of  H,  h1}  and 
hz  at  the  different  positions  of  the  weights.  With  these  values 
substituted  in  the  formula  (15),  the  r.  p.  m.  of  the  governor  is 
readily  determined.  The  governor  of  Fig.  101  is  the  one  commonly 
used  on  Corliss  engines. 

133.  The  Spring-balanced  Governor.  —  In  most  high-speed 
engines,  the  centrifugal  force  of  the  revolving  weight  is  balanced 


FIG.  102 


by  the  force  of  a  spring.  Figure  102  shows  a  weight  W,  revolv- 
ing about  the  center  of  a  spindle  or  shaft.  The  centrifugal  force 
C  of  the  weight  is  balanced  by  the  spring  tension  S.  The  weight 
is  at  a  distance  R  from  the  center  of  rotation.  Hence  we  have 


For  a  certain  value  of  n,  C  varies  directly  as  R.  In  Fig.  103, 
this  variation  is  shown  graphically.  At  speed  n  and  with  a 
radius  R,  the  centrifugal  force  is  C.  If  R  is  doubled,  C  is  doubled. 
For  other  values  of  n,  as  HI  and  WQ,  C  will  have  different  values 
with  the  same  radius  R. 


144 


ENGINES   AND    BOILERS 


In  any  kind  of  uniform  spring  the  elongation,  shortening,  or 
deflection  is  proportional  to  the  force  producing  the  deformation. 
Figure  104  shows  graphically  the  relation  between  the  pull  of  the 
spring  and  the  elongation.  If  Figs.  103  and  104  be  superimposed, 
as  in  Fig.  105,  we  easily  can  see  the  relations  that  must  exist  if 


of  p0/h  of 
FIG.  103 


o 

FIG.  104 


the  spring  pull  equals  the  centrifugal  force.  For  the  three  speeds 
HI,  n  and  n2,  the  spring  pull  equals  the  centrifugal  force,  as  seen 
at  the  points  b,  d,  and  /.  That  is,  the  spring  pull  will  balance 
the  centrifugal  force  at  the  speed  %when  the  elongation  of  the  spring 
is  ei.  The  weight  will  then  be  revolving  at  a  radius  Ri=a+e, 
where  a  is  the  distance  the  weight  would  be  from  the  center 

of  rotation  if  there  were  zero 
tension  in  the  spring.  At  a 
speed  n,  the  forces  will  balance 
when  the  elongation  is  e  and 
the  radius  is  R=a-\-e.  In  like 
manner,  the  forces  balance  at 
the  speed  n2  with  an  elongation 
62  and  radius  R%.  If  the  origin 
A  should  be  moved  over  to  the 
origin  0,  i.e.  the  distance  a  be 
made  zero,  it  is  seen  that  the 
spring  pull  could  balance  the 
centrifugal  force  at  only  one  speed.  This  means  that  the  gover- 
nor would  then  be  isochronous.  If  the  origin  A  were  moved  to 
the  left  of  the  origin  0,  the  governor  would  be  unstable,  because 
the  speed  HI  at  no  load  would  be  less  than  the  speed  712  at  full 
load.  As  stated  before,  an  unstable  or  isochronous  governor  could 
not  be  used  in  practice. 


-  Q  . 


FIG.  105 


GOVERNORS  145 

What  is  known  as  scale  of  spring  is  the  force  necessary  to  pro- 
duce an  elongation  of  one  inch  in  the  spring.  If  the  spring  pull 
is  Si  at  elongation  e\,  and  $2  at  elongation  e2,  the  scale  of  the 
spring  is  equal  to 


Suppose  that  we  desire  to  find  the  scale  of  spring  necessary  for 
a  governor  such  as  that  shown  in  Fig.  102.  Let  us  suppose  that 
the  no-load  speed  HI  and  the  full-load  speed  n2,  and  the  corre- 
sponding radii  RI  and  R2,  are  known.  First  compute  Ci  and  C2 
for  the  two  speeds.  Then,  since  the  spring  pull  must  equal  the 
centrifugal  force  at  all  loads,  the  scale  of  spring  is  seen  to  be 


because  R\  —  R^=ei  —  e%. 

In  actual  governors  it  is  very  seldom  that  the  spring  pull  acts 
in  the  same  line  as  the  centrifugal  force.     Figure  106  represents 


FIG.  106 

a  more  usual  case.  In  the  solution  of  this  problem,  moments  will 
be  taken  about  the  pivot  point  of  the  governor  arm  B.  In  the 
full-line  position,  the  moment  of  the  centrifugal  force  equals  the 
moment  of  the  spring  pull  about  the  point  B.  That  is,  CXh  = 
SXd.  In  like  manner,  CiXi=SiXf.  The  scale  of  spring  equals 
(S  —  Si)  divided  by  the  elongation  of  the  spring  as  the  weight 
goes  from  R  to  RI. 


146  ENGINES   AND    BOILERS 

134.  Governing  by  Changing  Position  of  Eccentric.  —  Most 
shaft  governors  regulate  the  steam  supply  by  changing  the  per- 
cent of  cut-off.  This  is  accomplished  by  changing  the  position 
of  the  eccentric  relative  to  the  crank.  In  Fig.  107,  a  Zeuner 
valve  diagram  is  shown  on  which  appear  two  positions  of  the 
crank  at  cut-off.  In  the  full-line  construction,  cut-off  comes  late 
and  the  angle  of  advance  is  «i,  i.e.  when  the  crank  is  on  head- 
end dead  center,  the  eccentric  will  be  at  E\  (direct  valve).  By 
shifting  the  eccentric  forward  to  E^  the  angle  of  advance  is 


FIG.  107  FIG.  108 

changed  to  0$.  The  dotted  construction  gives  the  position  of 
the  crank  with  the  new  angle  of  advance.  It  is  seen  that  the 
cut-off  comes  earlier  with  the  larger  angle  of  advance.  By  turning 
the  eccentric  on  the  shaft,  the  time  of  cut-off  can  be  changed  but 
the  value  of  the  steam  lap  will  be  the  same  for  all  values  of  a, 
because  the  only  way  to  change  the  laps  is  to  move  the  valve 
on  the  stem. 

Figure  108  shows  the  effect  on  cut-off  of  changing  the  valve 
travel  while  keeping  the  angle  of  advance  constant.  The  Zeuner 
construction  for  a  large  valve-travel  is  shown  in  full  line.  Cut- 
off is  seen  to  come  fairly  late.  With  a  smaller  valve-travel,  as 
shown  by  the  dotted  construction,  cut-off  is  seen  to  come  earlier. 
Hence  it  appears  that  late  cut-off  is  obtained  with  large  valve- 
travel  and  earlier  cut-off  with  small  valve-travel. 

A  governor  can  regulate  cut-off  by  changing  either  a  or  the 
valve-travel.  It  may  be  so  constructed  that  it  will  control  by 
making  only  the  one  change  or  it  may  change  the  valve-travel 
and  the  angle  of  advance  at  the  same  time. 


GOVERNORS 


147 


Changing  the  angle  of  advance  affects  the  other  events  as  well 
as  cut-off.  When  a  is  increased  all  events  occur  sooner.  Thus 
on  one-valve  engines  that  control  the  steam  supply  by  shifting 
the  eccentric,  it  is  found  that  a  high  compression  accompanies 
early  cut-off,  such  as  is  characteristic  of  the  Stephenson  valve-gear. 

135.  Governing  by  Changing  a.  —  In  Fig.  109,  a  governor  is 
shown  that  controls  cut-off  by  turning  the  eccentric  around  the 
shaft.  The  two  weight  arms  are  pivoted  at  the  points  G  and  F. 


FIG.  109 


Any  rotation  of  these  arms  about  their  pivots  causes  the  eccentric 
to  turn  on  the  shaft.  The  weight  arms  are  connected  by  the 
links  AC  and  BD  to  AB,  which  carries  the  eccentric.  At  light 
loads  the  speed  of  the  engine  is  greater  than  at  heavy  loads,  and 
the  weights  will  be  farther  from  the  center  of  rotation.  This 
movement  of  the  weights  away  from  the  center  of  rotation  causes 
the  eccentric  to  turn  in  a  clockwise  direction  relative  to  the 
crank.  It  is  evident  that  this  increases  a  arid  therefore  makes 
cut-off  come  earlier,  as  it  should  at  light  loads. 

136.   Governing  by  Changing  both  a  and  the  Valve-travel.  - 

In  Fig.  110  a  governor  is  shown  that  changes  both  a  and  the 
valve-travel  at  the  same  time.  The  pivot  point  on  the  flywheel 
carries  the  governor  arm.  The  eccentric  is  shown  as  a  pin. 
When  the  governor  arm  moves  about  the  pivot  point  in  a  clock- 
wise direction,  it  carries  the  center  of  the  eccentric  with  it  and 


148 


ENGINES   AND    BOILERS 


makes  a  smaller  and  the  valve-travel  larger.  It  is  known  that 
small  a  and  large  valve-travel  both  give  late  cut-off.  Conversely, 
a  counter-clockwise  movement  of  the  arm  about  the  pivot  point 
gives  early  cut-off  because  it  makes  a  larger  and  the  valve-travel 
smaller.  The  centrifugal  force  acts  through  the  center  of  rota- 
tion and  the  center  of  gravity.  Hence  this  force  tends  to  give 
the  arm  counter-clockwise  movement  about  the  pivot.  At  light 
loads,  and  therefore  higher  speeds,  the  centrifugal  force  will  be 
greater  than  for  heavy  loads.  This  tends  to  give  the  arm  counter- 
clockwise rotation  about  the  pivot  and  this  makes  cut-off  come 


FIG.  110 

earlier  for  light  loads,  as  it  should.  A  great  many  other  governors 
besides  the  one  shown  in  Fig.  110,  change  a  and  the  valve-travel 
in  the  same  way. 

137.  Centrifugal  and  Inertia  Governors.  —  As  has  been  previ- 
ously stated,  all  governor  weights  have  inertia.  If  the  tendency 
of  the  weight  to  keep  moving  at  the  same  speed  helps  to  effect 
the  change  of  position  that  causes  the  governing,  the  governor 
will  respond  more  quickly  than  it  would  if  the  inertia  opposed 
the  change.  In  the  gravity-balanced  spindle  governors  that  were 
considered  in  §132,  inertia  acts  against  the  rapid  change  of  posi- 
tion of  the  balls  and  so  tends  to  make  the  governor  sluggish. 

In  the  governor  of  Fig.  110,  the  inertia  of  the  arm  assists  in 
the  governing.  If  the  load  is  suddenly  thrown  off  the  engine, 


GOVERNORS  149 

it  will  momentarily  speed  up.  This  means  that  the  flywheel  will 
go  ahead  of  the  governor  arm  or  rotate  in  a  clockwise  direction 
relative  to  the  arm.  This  swings  the  eccentric  nearer  the  center, 
which  makes  a  large  and  the  valve-travel  small.  Hence  cut-off 
occurs  sooner,  which  tends  to  re-establish  the  conditions  of  equi- 
librium. In  Fig.  110  the  arm  is  made  very  heavy,  so  that  it  will 
have  considerable  inertia. 

It  must  not  be  assumed  that  the  inertia  of  the  arm  is  the  only 
factor  in  the  governing.  The  centrifugal  force  also  plays  its 
part,  as  has  been  explained  previously.  The  governor  of  Fig. 
110,  while  it  is  very  simple  in  construction,  is  at  the  same  time 
sensitive  and  quick  in  its  action.  It  is  widely  used  and  is  known 
as  the  Rites  inertia  governor. 


CHAPTER  X 
STEAM  TURBINES 

138.  Introduction. — The  steam  turbine  of  to-day  is  of  as  much 
importance  in  the  world  of  engineering  as  is  the  reciprocating 
steam  engine.     Practically  all  large  steam  power  plants  which 
produce  electric  current  employ  the  turbine  engine.     The  devel- 
opment of  the  turbine  has  been  remarkably  rapid.     However,  it 
would  not  be  true  to  say  that  the  turbine  has  crowded  the  recip- 
rocating engine  from  the  power-plant  field.     The  fact  is  rather 
that  it  has  developed  along  new  and  different  lines  of  use,  and 
now  occupies  a  field  that  was  never  held  by  the  reciprocating 
engine,  i.e.  as  the  direct  connected  prime  mover  for  high-speed 
electric  generating  units. 

The  turbine  and  the  alternating-current  generator  have  devel- 
oped together.  Both  the  turbine  and  the  alternating-current  gen- 
erator are  well  adapted  to  high-speed  rotation.  The  cost  of  a  slow- 
speed  electric  generator  is  much  higher  than  that  of  a  high-speed 
generator  of  the  same  capacity.  Before  the  days  of  the  turbine, 
generators  were  nearly  all  of  the  direct-current  type,  which  could 
not  run  at  very  high  speeds. 

The  electric  system  of  power  transmission  is  more  economical 
than  the  old  belt  system.  Hence  the  turbine  has  replaced  the 
reciprocating  engine  in  some  manufacturing  plants  on  account  of 
the  development  of  systems  of  electric  power  transmission.  On 
land,  large  turbines  are  seldom  used  to  drive  anything  but 
electric  generators. 

139.  History.  —  The  steam  turbine  is  not  a  modern  invention. 
Hundreds  of  years  ago  people  knew,  as  every  child  knows  today, 
that  a  pin-wheel  would  rotate  when  blown  upon.     There  are  rec- 
ords of  turbines  built  in  the  quite  distant  past.     They  were  little 
but  toys,  however,  like  pin-wheels,  and  of  no  practical  importance. 
The  modern  turbine  dates  from  the  years  between  1880  and  1890. 
During  this  period  two  types  of  turbines  that  have  become  of 
great  practical  importance  were  developed. 

DE  LAVAL,  the  inventor  of  the  cream  separator,  sought  to  drive 
his  separator  by  means  of  a  turbine.  After  several  experiments, 
he  perfected  a  type  that  was  satisfactory  for  that  purpose.  The 

150 


STEAM   TURBINES  151 

same  turbine,  with  improvements,  has  been  used  in  large  numbers 
for  driving  centrifugal  pumps,  fans,  and  even  small  generators. 

During  practically  the  same  time,  C.  A.  PARSONS  developed  the 
type  of  turbine  that  now  bears  his  name.  These  two  pioneers 
were  soon  followed  by  other  experimenters.  Various  forms  of  tur- 
bines were  developed.  Some  of  these  are  still  used.  Others  are 
obsolete  and  are  of  interest  only  from  an  historical  standpoint. 

In  the  development  of  the  turbine,  there  were  two  obstacles 
that  had  to  be  overcome.  The  first  of  these  was  the  lack  of 
knowledge  of  principles.  The  second  was  the  need  of  better 
mechanical  means  of  manufacture.  As  will  be  shown  later,  the 
velocity  of  the  rotor  in  a  turbine  must  be  very  high.  This  causes 
large  stresses,  and  makes  necessary  a  very  perfect  balance.  The 
clearances  between  the  rotors  and  the  stationary  parts  must  be 
small  to  prevent  undue  leakage.  This  calls  for  an  excellence  of 
design  and  construction  that  did  not  commonly  exist  for  heavy 
machines  in  the  past.  As  in  the  development  of  any  new  machine, 
satisfactory  solutions  of  the  problems  grew  out  of  the  necessities, 
so  that  the  modern  turbine  is  as  reliable  and  dependable  as  any 
piece  of  machinery  in  the  power  plant. 

140.  Fundamental  Principles.  —  Before  making  a  study  of  the 
common  types  of  turbines  now  in  use,  we  shall  discuss  the  fun- 
damental principles  of  the  steam  turbine.  It  is  not  our  purpose 
to  give  an  exhaustive  discussion,  but  only  to  present  the  principles 
in  their  simplest  form.  The  sketches  of  blades  and  nozzles  are 
not  exactly  correct  in  shape  for  the  conditions  assumed.  They  are 
to  be  considered  as  only  diagrammatic. 

Steam  under  pressure  contains  a  certain  amount  of  usable  heat. 
The  available  amount  depends  upon  the  initial  pressure,  the 
degree  of  superheat,  and  the  pressure  to  which  the  steam  may 
be  dropped.  There  is  the  same  amount  of  available  heat  if  the 
initial  and  final  conditions  of  the  steam  are  the  same,  no  matter 
whether  we  are  considering  the  reciprocating  steam  engine  or  the 
steam  turbine.  The  turbine,  or  the  reciprocating  engine,  is  effi- 
cient, or  is  not,  according  as  it  uses  a  large  or  a  small  amount 
of  this  available  energy. 

Since  the  turbine  and  the  reciprocating  engine  both  use  the 
same  medium,  it  is  not  to  be  expected  that  one  will  be  much 
more  efficient  than  the  other.  Both  reciprocating  engines  and 


152  ENGINES   AND   BOILERS 

turbines  may  be  made  of  about  the  same  thermal  efficiency.  The 
choice  of  type  of  engine  depends  upon  other  considerations  than 
efficiency. 

The  greatest  loss  in  the  reciprocating  engine  is  due  to  the  initial 
condensation  in  the  cylinder.  Since  the  cylinder  walls  are  made 
of  a  heat-conducting  material,  they  will  never  be  as  hot  as  the 
incoming  high-pressure  steam,  and  they  will  be  hotter  than  the 
low-pressure  steam  leaving  the  cylinder.  The  relatively  hot 
steam  coming  to  the  cylinder  strikes  the  cooler  cylinder  walls 
and  some  condensation  takes  place,  with  a  consequent  shrinkage 
in  the  volume.  The  condensed  steam  is  mostly  re-evaporated 
before  the  steam  leaves  the  cylinder,  owing  to  an  absorption  of  heat 
from  the  then  hotter  cylinder  walls. 

The  loss  in  the  turbine  is  due  to  other  causes,  such  as  leakage, 
friction,  etc.  The  leakage  occurs  around  the  ends  of  the  blades 
or  from  stage  to  stage.  The  friction  exists  between  the  steam  and 
the  parts  of  the  turbine  in  the  passage  of  steam  through  both 
the  stationary  and  the  moving  parts.  There  is  also  a  windage 
loss  between  the  moving  parts  and  the  steam.  This  friction  does 
not  cause  a  complete  loss,  because  a  part  of  the  heat  generated 
may  be  used  in  later  stages  of  the  turbine. 

141.  Available  Energy  in  Steam.  —  In  order  to  make  clear  the 
nature  of  the  available  energy  in  steam,  a  concrete  example  will 
be  taken. 

(1)  Let  us  assume  that  the  steam  is  dry  saturated  steam  at  a 
pressure  of  150  pounds  gage  (165  pounds  absolute),  and  that  it 
is  allowed  to  expand  adiabatically  to  a  pressure  of  15  pounds 
absolute.*  The  heat  contents  of  a  pound  of  dry  saturated  steam 
at  165  pounds  absolute  is  1194  B.t.u.  The  heat  contents  of  a 
pound  of  steam  at  15  pounds  absolute,  after  expanding  adiabati- 
cally from  165  pounds,  is  1019  B.t.u.  The  difference  between 
these  values,  which  is  the  amount  of  heat  available  for  doing 
work,  is  1194—1019  =  175  B.t.u.  At  165  pounds  pressure,  dry 
saturated  steam  occupies  a  volume  of  2.75  cubic  feet  per  pound. 
At  15  pounds  pressure,  after  the  expansion  just  mentioned,  the 
quality  is  87  per  cent,  and  the  volume  is  about  23  cubic  feet. 

In  the  reciprocating  steam  engine,  this  change  in  volume,  work- 
ing by  its  pressure,  does  work  on  the  piston  in  forcing  it  forward. 

*  Adiabatic  expansion  is  that  in  which  no  heat  is  added  to  the  steam  and  none  is  extracted 
except  by  the  conversion  of  heat  into  work. 


STEAM   TURBINES  153 

The  velocity  of  the  piston  is  immaterial.  In  the  turbine,  the 
same  change  in  volume  takes  place,  but  the  steam  is  allowed  to 
acquire  velocity  in  expanding.  The  energy  of  the  steam  due  to 
its  velocity  is  imparted  to  the  rotor  of  the  turbine.  The  efficiency 
of  a  perfect  engine  working  on  the  Rankine  cycle,*  between  the 
pressures  of  165  and  15  pounds  absolute,  is  175/(1194— 181)  =17 
per  cent.  Since  neither  the  reciprocating  engine  nor  the  turbine 
is  perfect,  neither  would  have  an  efficiency  as  great  as  17  per 
cent  when  worked  between  the  pressure  limits  named. 

(2)  Assume  that  the  steam  is  allowed  to  expand  adiabatically 
from  165  pounds  absolute  to  1  pound  absolute.  The  heat-drop 
is  1194  —  871  =  323  B.t.u.  and  the  efficiency  on  the  Rankine  cycle 
is  323/(1194-70)  =  28.6  per  cent. 

142.  Velocity  Due  to  Expansion.  —  Let  us  next  compute  the 
velocity  of  the  steam  if  all  the  heat-drop  goes  to  giving  the  steam 
velocity. 

(1)  Suppose  that  the  drop  in  pressure  is  from  165  to  15  pounds 
absolute.  Since  one  B.t.u.  =  778  foot-pounds,  the  energy  is  175  X 
778  =  136,100  foot-pounds  per  pound  of  steam.  The  energy  of 
motion,  or  kinetic  energy,  is  mv2/2  =  (1/32)  Xv2/2.  This  must  be 
equal  to  the  value  136,100  foot-pounds  just  calculated.  Hence 

v2  =  64X136,100  =  8,710,400, 
v  =  2950  ft./sec. 

(2)  If  the  drop  in  pressure  is  from  165  to  1  pound,  we  find,  in  a 
similar  manner, 

K.E.  =  778X323  =  (1/32)  Xv*/2, 
whence 

v  =  4010  ft./sec. 

In  the  steam  turbine,  the  steam  must  be  expanded,  and  the 
velocity  due  to  this  expansion  must  be  used  by  imparting  its 
kinetic  energy  to  the  rotor  of  the  turbine.  If  this  is  done  by 
allowing  the  steam  to  expand  in  the  stationary  parts  of  the  tur- 
bine and  imparting  the  velocity  thus  produced  to  the  moving 
parts,  the  turbine  is  said  to  be  of  the  impulse  type.  If  it  is  done 

*  To  compare  steam  engines,  the  efficiencies  based  on  the  Rankine  cycle  are  often  used.  The 
efficiency  of  a  steam  engine  operating  on  the  Rankine  cycle  is  given  by  the  expression 
(Qi  —  Qi/Q\,  where  Q\  is  the  amount  of  heat  required  to  make  dry  steam  at  boiler  pressure  from 
water  at  the  temperature  of  the  exhaust,  and  Qz  is  the  amount  of  heat  rejected  from  the  engine 
minus  the  heat  of  the  liquid  at  the  temperature  of  the  exhaust. 


154  ENGINES   AND   BOILERS 

by  allowing  the  steam  to  expand  in  the  moving  parts,  the  unbal- 
anced steam  pressure  reacting  on  the  rotor,  the  turbine  is 
called  a  reaction  turbine.  In  some  turbines  the  steam  expands 
both  in  the  stationary  parts  and  in  moving  parts,  and  the  turbine 
is  said  to  be  of  the  impure  reaction  type. 

143.  Impulse  and  Reaction.  —  In  order  to  understand  the  prin- 
ciples involved,  consider  the  simplest  cases  involving  the  principles 
of  impulse  and  reaction  in  which  the  velocity  is  created  and  used. 
It  may  be  easier  to  think  of  the  jet  as  a  jet  of  water,  for  in  that 
case  the  fluid  does  not  expand  when  the  pressure  on  it  is  reduced. 

Otherwise,  the  steam  jet  and  the 

""""*'*•         .        e'^    *      :    i    water  jet  follow  the  same  laws  of 


impulse  and  reaction. 
FlG  jjj  Suppose  that  water  issues  from 

a   nozzle,  as    in  Fig.   111.      The 

nozzle  is  stationary,  and  the  issuing  jet  has  a  velocity  of  v  feet 
per  second.  The  unit  mass  m  of  water  that  will  be  considered  is 
that  issuing  from  the  nozzle  in  one  second.  A  particle  of  water 
in  the  jet  will  move  v  feet  in  one  second.  The  kinetic  energy 
of  this  unit  mass  will  be  mv2/2. 

(1)  Consider  the  case  in  which  the  jet  strikes  a  stationary  flat 
surface  (Fig.  112).     After  striking  the  flat  surface,  the  water  flows 
or  splatters  out  to  the  sides  at  right  angles  to  its  former  direction 
of  motion,  that  is  it  loses  all  its 

velocity  in  the  direction  of  the 

jet.    The  force  exerted  by  the     |  ^"," "'      *    ^  * 

jet  on  the  flat  surface  may  be 

measured  by  the  force  F  neces- 

sary  to  hold  the  flat  surface 

stationary.     Since  force  =  mass  X  change  in  velocity  per  second, 

and  since  the  time  in  which  mass  m  emerges  from  the  nozzle 

and  strikes  the  plate  is  one  second,  we  have  F  =  mv.     The  force 

exists  in  the  case  of  the  stationary  plate,  but  no  work  is  done 

because  the  plate  does  not  move. 

(2)  Suppose  that  the  flat  surface  moves  with  a  velocity  V  (Fig. 
112).     Then  the  force  F=mX(v-V).     The  quantity  (v-V)  is 
the  velocity  of  the  jet  relative  to  the  flat  surface.     It  is  seen  that 
F  is  less  than  before,  and  will  be  zero  if  the  velocity  of  the  sur- 
face is  the  same  as  the  velocity  of  the  jet.    The  work  done  in  one 


Force  F 

*  t/t/.  &    „ 


STEAM   TURBINES  155 

second  by  the  jet  on  the  plate  equals  F  times  the  distance  the 
plate  moves,  or 

W=FxV=m(v-V)V. 

The  velocity  of  the  plate  at  which  the  work  is  a  maximum  may 
be  found  by  equating  to  zero  the  first  derivative  of  the  work 
with  respect  to  V.  This  gives 


Hence  the  maximum  work  occurs  when  V  =v/2,  that  is,  when 
the  velocity  of  the  plate  is  half  that  of  the  jet. 

(3)  If  instead  of  striking  a  flat  surface,  the  jet  strikes  a  station- 
ary curved  surface,  such  that 

the  jet  is  turned  completely 
back  on  itself,  or  through  an 
angle  of  180°  (Fig.  113),  the 
force  F  exerted  on  the  surface 

is  mX2v,  which  is  twice  that  "*''  ^~*f 

,,       a   ,          f  PIG.  113 

exerted   on   the   nat   surtace. 

Since  the  curved  surface  is  stationary,  there  is  no  work  done. 

(4)  Suppose  the  curved  surface  (Fig.  113)  is  moving  with  a 
velocity  V.     The  velocity  of  the  jet  relative  to  the  curved  surface  is 
(v—V).     It  follows  that  the  absolute  velocity  of  the  jet  leaving 
the  surface  is  (v  —  V)  —  V  =  (v  —  2  V  )  .     Consequently  the  change  in 
the   velocity   of   the   jet   is    v-\-(v  —  2V)  =  2(v  —  V),    because    the 
direction  of  motion  is  completely  reversed.     As  before,  we  have 

F=mX2(v-V), 
and  the  work  done  per  second  is 

W=FV  =  2m(v-V)V  =  2m(vV-V*), 
whence 


For  maximum  work  we  must  have 

^=0 

dV 

Hence 

v=2V,  or  V=v/2. 

That  is  to  say,  the  curved  surface  should  move  at  half  the  velocity 
of  the  jet  for  the  production  of  maximum  work.     If  the  latter  condi- 


156  ENGINES   AND   BOILERS 

tion  exists,  the  absolute  velocity  of  the  jet  as  it  leaves  the  sur- 
face is  zero.  That  is,  all  of  the  velocity  of  the  jet  has  been  used. 
(5)  In  Fig.  114  we  have  a  tank  that  is  free  to  move  horizontally 
upon  a  track.  In  one  side  of  this  tank  is  placed  an  orifice  or 
nozzle.  The  water  issues  from  this  nozzle  due  to  the  pressure  of 
the  water  from  above.  If  the  tank  is  stationary,  the  water  leaves 
the  tank  with  an  absolute  velocity  v.  The  force  F,  due  to  the 

unbalanced  pressure  of  the  water 
in  the  tank,  tends  to   force   the 
tank  to  the  left,  but  since  the  tank 
|  is    held    stationary,   the    force  F 

^:   fe/.v-^       does   no   work.      If   the    tank   is 
'       allowed  to  move  to  the  left  with 

.....,, ,,,JJj$,,,,,,^),f,,,,,,,,,,,r          a  vel°city  V,  however,  the  work 

piG  114  done  will  be  FV.     The  absolute 

velocity  of  water  leaving  the  tank 

is  (v—V).     The  maximum  work  will  be  done  when  V  =v,  that  is 
when  the  escaping  water  has  no  absolute  velocity. 

In  the  first  four  cases  considered,  the  jet  impinged  on  a  sur- 
face. Work  was  done  by  the  jet  striking  and  moving  the  surface. 
A  turbine  in  which  the  pressure-drop  occurs  in  a  stationary  nozzle 
or  part  is  said  to  be  an  impulse  turbine  (§142)  because  the  energy 
is  given  to  the  moving  parts  by  the  impulse  of  the  jet.  In  the 
fifth  case  the  drop  in  pressure  occurred  in  the  moving  nozzle. 
When  this  occurs  in  a  turbine,  it  is  said  to  be  a  reaction  turbine 
(§142).  A  comparison  of  the  above  simple  examples  shows  that 
the  velocity  of  the  moving  parts  of  a  reaction  turbine  must  be 
nearly  twice  as  great  as  that  of  the  impulse  type,  other  factors 
being  equal. 

144.  Bucket  Shapes.  —  In  the  common  types  of  steam  tur- 
bines, buckets  or  blades  are  mounted  on  the  periphery  of  a  wheel 
or  rotor.  The  shape  of  these  blades  is  something  like  that  of 
the  curved  surface  considered  in  (3),  §143.  Of  necessity  the  jet 
cannot  be  completely  turned  through  an  angle  of  180°  as  in 
(3),  §143,  because  the  steam  must  have  a  velocity  in  the  direc- 
tion of  the  axis  of  rotation  of  the  rotor  in  order  to  get  to  the 
bucket  and  to  leave  it. 

In  Fig.  115,  let  Vi  denote  the  velocity  of  the  jet  relative  to  the 
bucket  or  blade  at  the  point  where  the  jet  first  strikes  it,  and  let  a 


STEAM   TURBINES 


157 


denote  the  angle  it  makes  with  the  tangent  to  the  rim  of  the  rotor. 
Let  v%  and  0,  respectively,  denote  the  velocity  and  the  angle  upon 
leaving  the  bucket.  We  see  that  the  component  of  the  velocity 
of  the  jet  relative  to  the  bucket  in  the  direction  of  the  tangent 
is  vi  cos  a  and  the  component  in  the  direction  of  the  axis  of  the 
rotor  is  v\  sin  a.  In  like  man- 
ner, the  relative  velocity  v^ 
has  similar  components  v%  cos  /3 
and  t>2  sin  /3.  These  compo- 
nents Vi  sin  a  and  v%  sin  /3 
must  be  large  enough  to  get 
the  jet  through  the  row  of 
buckets  on  the  rotor,  in  order 
that  the  following  buckets 
shall  not  interfere  with  the 
flow. 

If  the  relative  velocity  of  FIG.  115 

the  jet  is  Vi  and  the  angle  that 

it  makes  with  the  tangent  is  a,  the  absolute  velocity  v  of  the 
jet  makes  a  different  angle  6  with  the  tangent  (Fig.  116).  If  V 
denotes  the  tangential  velocity  of  the  bucket,  v  is  the  resultant  of 
the  two  velocities  Vi  and  V,  and  6  is  the  angle  that  this  resultant 
v  makes  with  the  tangent. 

In  like  manner,  the  resultant  of  v%  and  V  at  the  exit  is  the  abso- 
lute velocity  v'  at  the  exit,  and 
it  makes  an  angle  <f>  with  the 
tangent. 

If  we  assume  that  the  jet 
strikes  the  bucket  in  the  direc- 
tion of  the  tangent  to  the  rim 
of  the  rotor,  the  preceding  par- 
agraph shows  that  an  error  will 
be  introduced.  In  order  to 
make  the  calculations  as  simple 
as  possible,  however,  we  shall 
assume  that  the  jet  does  strike 
tangentially,  and  we  shall  bear  in  mind  that  some  error  has 
been  introduced.  The  results  previously  derived  for  steam  veloci- 
ties for  certain  heat  drops  will  now  be  applied  to  the  problem 
of  the  turbine. 


FIG.  116 


158  ENGINES   AND   BOILERS 

145.  The  Single-stage  Turbine.  —  In  a  single-stage  impulse 
turbine,  the  steam  is  expanded  in  a  stationary  nozzle,  and  is 
directed  against  the  moving  buckets  or  blades,  which  are  mounted 
on  the  rim  of  the  rotor.  In  the  single-stage  reaction  turbine,  the 
rotor  itself  carries  the  nozzles,  and  the  steam  expands  in  passing 
through  them. 

If  the  expansion  of  the  steam  is  from  165  to  15  pounds  absolute 
we  have  seen  in  (1),  §  142,  that  the  steam  or  jet  velocity  is  2950 
feet  per  second.  For  maximum  work  done,  the  bucket  velocity 
of  the  impulse  turbine  is  half  that  of  the  jet  velocity  (§  143). 
Hence  the  peripheral  velocity  of  the  rotor  should  be  2950/2  =  1475 
feet  per  second.  With  a  reaction  turbine,  the  peripheral  velocity 
of  the  rotor  should  be  the  same  as  the  jet  velocity,  or  2950  feet 
per  second. 

If  the  rotor  speed  is  assumed  to  be  3000  revolutions  per  minute, 
or  50  revolutions  per  second,  the  diameter  of  the  rotor  should  be 

1475 

-^r-  =9.4  feet 

50-7T 

for  an  impulse  wheel.  This  is  obviously  very  much  too  large. 
For  a  reaction  wheel,  the  diameter  should  be  18.7  feet,  which  is 
absurd.  If  a  speed  of  24,000  r.  p.  m.  is  assumed,  the  diameter 
of  the  rotor  for  an  impulse  turbine  should  be 

1475 


or  14  inches.  These  values  for  the  speed  and  the  diameter  of 
the  rotor  are  not  far  from  those  which  are  used  in  the  DeLaval 
single-stage  turbine. 

The  preceding  examples  show  what  a  very  high  peripheral  veloc- 
ity is  necessary  for  a  fair  efficiency  in  a  single-stage  turbine. 
With  a  vacuum,  a  much  larger  velocity  should  be  used.  Since 
immense  stresses  are  induced  in  the  wheel  by  these  high  velocities, 
it  is  readily  seen  that  a  single-stage  reaction  turbine  is  almost  out 
of  the  question.  If  such  turbines  were  operated,  their  efficiency 
would  necessarily  be  very  low.  Hence  they  are  not  used. 

In  Fig.  117,  a  diagram  of  the  single-stage  impulse  turbine  is 
shown.  Steam  enters  the  nozzle  from  the  left,  expanding  as 
it  passes  through.  As  the  pressure  drops,  a  high  velocity  is 
imparted  to  the  steam.  The  steam  leaves  the  nozzle  at  low  pres- 
sure and  at  a  high  velocity.  The  steam  now  impinges  upon  the 


STEAM   TURBINES 


159 


buckets  or  blades  of  the  rotor,  imparting  to  the  rotor  its  velocity, 
and  therefore  its  kinetic  energy.  Upon  leaving  the  rotor,  the 
absolute  velocity  of  the  steam  is  quite  low. 

The  graphs  at  the  lower  part  of  the  diagram  show  the  changes 
in  the  pressure  and  in  the  velocity.  The  steam  pressure  is  shown 
by  the  full  line,  and  the  steam  velocity  by  the  dotted  line.  While 
single-stage  impulse  turbines  are  widely  used,  they  are  never 
made  in  large  sizes. 

The  diagram  of  Fig.  118  represents  a  single-stage  reaction  tur- 
bine. The  steam  passes  from  the  left  directly  to  the  rotor.  The 
rotor  carries  blades  so  shaped  that  the  spaces  between  them  act 
as  nozzles.  The  steam  expands  in  these  spaces  or  nozzles.  As 
it  expands,  its  pressure  drops,  and  it  reacts  upon  the  blades.  This 


Single-Stage  Impulse 
FIG.  117 


Single  -  Sfagre    Reac  tioh 
FIG.  118 


force  of  the  steam  on  the  blades  causes  the  rotor  to  move  and  to 
absorb  the  energy  liberated  by  the  expansion.  It  will  be  noticed 
from  the  graphs  that  the  velocity  does  not  change  much  in  pass- 
ing through  the  rotor,  but  the  pressure  drops  during  the  passage. 
In  order  to  decrease  the  peripheral  velocity  of  the  rotor,  and 
at  the  same  time  to  expand  and  use  all  the  velocity  of  the  steam, 
more  than  one  set  of  rotor  blades  or  buckets  are  employed.  This 
is  called  staging.  The  steam  passes  successively  through  the 
sets  of  blades  in  each  stage,  giving  up  part  of  the  energy  to  each  set. 


160 


ENGINES   AND    BOILERS 


Staging.  —  In  multi-stage  impulse  turbines,  two  methods 
are  in  use.  The  first  method  is  to  expand  the  steam  in  one  set 
of  stationary  nozzles,  and  to  take  out  part  of  the  velocity  in 
each  stage.  This  is  known  as  velocity-staging. 

The  second  method  is  to  expand  the  steam  partially  in  one  set  of 
stationary  nozzles,  using  up  the  velocity  caused  by  this  expan- 
sion in  one  stage,  then  to  expand  the  steam  again  in  another  set  of 
stationary  nozzles,  using  the  velocity  thus  generated  in  the 
second  stage,  and  so  on.  This  scheme  is  called  pressure-staging. 
A  combination  of  pressure-staging  and  velocity-staging  is  also 
used,  in  which  there  are  two  or  more  velocity  stages  in  each  pres- 
sure stage. 

147.  Multi-stage  Impulse  Type  with  Velocity-staging.  —  The 
diagram  in  Fig.  119  shows  the  velocity-stage  impulse  turbine. 


-  Velocit  y  -  Stage    /mpulse . 

FIG.  119 


The  steam  enters  from  the  left  and  passes  through  the  stationary 
expanding  nozzle,  where  the  pressure  drops  and  the  velocity  is 
acquired  in  exactly  the  same  manner  as  in  the  single-stage  im- 
pulse turbine  of  Fig.  117.  The  rotor  in  this  case,  however,  has 
much  less  velocity  than  the  rotor  shown  in  Fig.  117.  Hence  the 
steam  loses  only  a  part  of  its  velocity  in  passing  through  the  first 


STEAM   TURBINES  161 

set  of  buckets.  Emerging  from  the  first  set  of  buckets,  it  passes 
through  a  set  of  stationary  blades  or  vanes  which  change  the 
direction  of  flow  of  the  steam,  but  not  its  velocity.  These  sta- 
tionary blades  are  necessary,  because  the  steam  has  a  large  up- 
ward component  of  velocity  after  leaving  the  rotating  buckets 
of  the  first  stage;  and  since  the  velocity  of  the  rotor  is  downward, 
the  direction  of  flow  must  be  reversed  so  that  the  steam  may 
impinge  on  the  second  set  of  rotating  buckets.  In  passing  through 
the  second  set  of  moving  buckets,  more  of  the  steam  velocity  is 
taken  up  by  the  rotor.  The  direction  of  flow  is  again  changed 
by  the  stator,  and  so  on,  till  the  steam  finally  emerges  from  the 
turbine  with  its  velocity  practically  all  expended. 

Suppose,  for  example,  that  the  downward  velocity  of  the  steam 
as  it  leaves  the  nozzle  is  4000  feet  per  second,  and  that  the  bucket 
velocity  downward  is  500  feet  per  second.  As  it  leaves  the  first 
set  of  buckets  on  the  rotor,  the  steam  will  have  an  upward  velocity 
of  4000-2X500=3000  feet  per  second,  the  effect  of  friction 
being  neglected.  In  going  through  the  first  set  of  stator  blades, 
the  direction  of  flow  is  reversed,  but  is  unchanged  in  magnitude. 
Upon  leaving  the  second  set  of  rotor  buckets  its  velocity  will  be 
upward,  and  its  magnitude  will  be  3000-2X500=2000  feet  per 
second,  and  so  on  for  the  other  two  stages.  Each  set  of  moving 
buckets  takes  out  1000  feet  per  second  of  its  velocity,  and  it 
emerges  with  no  vertical  component  of  velocity. 

It  was  shown  in  §145  that  the  rotor  of  a  single-stage  turbine 
has  to  have  an  absurdly  large  diameter  unless  it  has  a  very  high 
speed  or  the  efficiency  is  very  low.  The  objection  to  such  high 
speed  is  that  the  turbine  must  have  a  reducing  gear  in  order  that 
the  power  may  be  used.  With  a  multi-stage  turbine  we  can 
choose  the  diameter  and  also  the  speed,  and  make  the  number  of 
stages  such  that  all  the  velocity  can  be  used. 

Let  us  assume  that  the  speed  is  3000  r.  p.  m.,  and  that  the 
diameter  of  the  rotor  is  3  feet.  Then  the  peripheral  velocity 
must  be  (3000/60)  X37r  =471  feet  per  second.  Neglecting  the 
effect  of  friction,  each  stage  will  absorb  a  steam  velocity  of  twice 
the  bucket  velocity,  or  942  feet  per  second.  If  the  steam  expands 
from  165  to  15  pounds  absolute,  the  steam  velocity  is  2950  feet 
per  second.  Hence  the  number  of  stages  necessary  will  be 
2950/942  =3.1,  and  three  stages  should  be  used.  If  the  turbine  is 
condensing,  and  the  pressure  drops  from  165  to  1  pound  absolute, 


162  ENGINES   AND    BOILERS 

the  steam  velocity  will  be  4010  feet  per  second,  the  number  of 
stages  4010/942  =4.2,  and  four  stages  should  be  used. 

In  a  velocity-stage  turbine,  the  efficiency  is  very  low  after  the 
first  two  stages,  principally  because  the  jet  is  broken  up  by  its 
passage  through  the  blades.  As  a  result,  more  than  two  velocity 
stages  are  seldom  used.  It  must  be  remembered  also  that  a  very 
high  steam  velocity  produces  a  very  great  friction  between  the 
steam  and  the  surfaces  of  the  blades,  thereby  causing  a  consid- 
erable loss. 


Impulse  Type  with  Pressure-staging.  —  If, 

instead  of  expanding  the  steam  completely  in  one  nozzle,  we  ex- 
pand it  only  a  little  in  the  first  nozzle,  then  use  its  velocity,  ex- 
pand it  some  more  in  a  second  nozzle,  and  again  use  the  velocity 
generated,  and  so  on,  the  process  is  called  pressure-  staging  (§146). 
Figure  120  shows  the  method  diagrammatically.  In  this  diagram 
there  are  five  sets  of  nozzles  and  five  pressure  stages.  The  steam 
enters  from  the  left  and  passes  through  the  stationary  nozzle. 
The  pressure  and  velocity  lines  below  show  that  the  drop  in  pres- 
sure is  accompanied  by  an  increase  in  velocity.  The  steam  with 
its  acquired  velocity  impinges  on  the  blades  of  the  rotor.  The 
velocity  is  absorbed  in  the  rotor.  As  the  steam  leaves  this  rotor 
with  low  velocity,  it  is  collected  and  led  to  a  second  stationary 
nozzle  in  which  the  pressure  is  again  dropped,  and  velocity  is 
acquired.  The  second  rotor  absorbs  this  velocity  and  the  steam 
passes  on  through  the  following  stages,  until  its  pressure  and 
velocity  are  practically  all  used  up  at  its  exit. 

Making  the  same  assumptions  for  speed  and  diameter  of  the 
rotor  as  in  the  preceding  type,  let  us  compute  the  number  of 
stages  necessary  with  the  pressure-stage  type.  If  the  diameter 
of  the  rotor  is  3  feet  and  the  speed  is  3000  r.  p.  m.,  there  is  a 
steam  velocity  of  942  feet  per  second  to  be  absorbed  per  stage 
(§147).  If  a  pound  of  steam  loses  a  velocity  of  942  feet  per 
second,  it  gives  up 


~XrX942  =  13900  foot-pounds 

Z       oZ 

of  kinetic  energy.     This  is  equivalent  to  13,900/778  =  17.9  B.t.u. 
For  the  non-condensing  condition  assumed  in  §147,  there  was  a 


^     «•     . 

STEAM   TURBINES 


163 


heat-drop  of  175  B.t.u.  available  in  the  whole  turbine.     The  num- 
ber of  stages  will  then  be 


For  the  condensing  conditions  assumed  in  §147,  the  number  of 
stages  will  be 


17.9" 

In  making  a  comparison  of  this  type  with  that  of  §147,  it 
might  seem  at  first  sight  that  the  velocity-staging  were  the  better, 
as  the  number  of  stages  is  so  much  smaller.  While  this  is  an 
advantage,  it  is  overbalanced  by  the  fact  that  the  pressure-stage 
type  is  more  efficient.  This  type  is  used  very  extensively  in 


Multi-  Pressure -Stage      Impulse. 

FIG.  120 

medium-sized  turbines.  It  is  often  called  the  multi-cellular  type 
because  each  pressure  stage  is  composed  of  a  cell  that  is  steam- 
tight  except  for  the  openings  through  the  inlet  and  outlet  nozzles. 
It  is  evident  that  it  is  necessary  to  keep  each  cell  steam-tight  in 
order  to  prevent  a  leakage  of  steam  from  one  stage  to  the  next. 
Where  no  difference  in  pressure  exists,  there  is  no  tendency  to 
leak.  In  the  velocity-stage  type,  the  pressure  was  the  same 
throughout  the  whole  turbine  beyond  the  expanding  nozzles,  and 
so  there  could  be  no  leakage. 


164 


'; 


ENGINES   AND   BOILERS 


149.  Multi-stage  Impulse  Type  with  Combined  Pressure- 
staging  and  Velocity-staging.  —  Very  often  a  combination  of 
pressure-staging  and  velocity-staging  is  used,  with  the  result  that 
some  of  the  advantages  of  both  types  are  utilized.  The  diagram 
of  Fig.  121  shows  this  arrangement.  In  the  sketch  three  pressure 
stages  are  shown,  and  each  pressure  stage  has  two  velocity  stages. 
The  drop  in  pressure  occurs  in  the  three  stationary  nozzles.  After 
expanding  in  the  nozzle,  the  steam  passes  through  the  buckets 


/*»"  Pressure 


g  &  Pressure  5fe?e  3  &  Pressure  5hj<?e 


Multi- Pressure,  Multi -Velocity  Stage  (Cvrf/sJ 
FIG.  121 

of  the  rotor  wheel,  where  a  part  of  the  velocity  is  absorbed.  It 
then  emerges  from  this  rotor,  and  its  direction  of  flow  is  reversed 
in  the  stationary  vanes,  as  in  the  velocity-stage  impulse  type. 
It  then  passes  through  a  second  set  of  moving  buckets,  where 
most  of  the  remaining  velocity  is  absorbed.  The  steam  is  now 
collected  and  expanded  some  more  in  the  next  nozzle,  and  the 
process  of  the  first  pressure  stage  is  repeated. 

By  dropping  the  pressure  in  three  successive  nozzles,  the  velocity 
generated  is  not  nearly  so  great  as  it  is  in  the  velocity-stage  type 
of  Fig.  119.  This  means  that  there  will  be  less  friction  loss  due 
to  excessively  high  steam  velocities.  Ordinarily  two  velocity 


STEAM   TURBINES  165 

stages  are  used  for  each  pressure  stage,  so  that  the  drop  in 
efficiency  due  to  a  disturbance  of  the  jet  is  not  so  great  as  in  the 
pure  velocity-stage  type. 

Applying  the  above  problem  to  this  combination  type,  let  us 
determine  the  number  of  pressure  stages  required.  Assuming,  as 
before,  that  the  speed  is  3000  r.  p.  m.,  and  that  the  diameter  is 
3  feet,  we  find  that  the  bucket  velocity  is  471  feet  per  second 
(§147).  Since  there  are  two  velocity  stages  for  each  pressure 
stage,  it  is  seen  that  2X2X471=1884  feet  per  second  of  steam 
velocity  per  pressure  stage  can  be  absorbed.  The  heat-drop  per 
pound  of  steam  that  corresponds  to  this  velocity  is 


yX—X  1884xl/778=  71.3  B.t.u. 

With  a  pressure-drop  from  165  to  1  pound  absolute,  in  which 
there  are  available  323  B.t.u.  for  doing  work,  there  will  be 
323/71.3  =  4  pressure  stages. 

Comparing  this  with  other  types,  it  is  seen  that  the  number 
of  pressure  stages  is  the  same  as  the  number  of  velocity  stages 
in  the  velocity-stage  type.  But  as  there  are  two  rows  of  rotor 
buckets  in  each  pressure  stage,  there  will  be  twice  as  many  rotor 
wheels  as  in  the  former  type.  With  pressure-staging,  there  were 
18  rows  of  rotor  buckets,  or  more  than  twice  as  many  as  in  the 
mixed  type.  This  combination  type  is  more  efficient  than  the 
velocity-stage  type,  and  at  the  same  time  it  is  more  compact  than 
the  pure  pressure-stage  type.  Due  to  these  distinct  advantages, 
the  combination  type  is  extensively  used. 

150.  Multi-stage  Reaction  Type.  —  In  the  pure  reaction  tur- 
bine, all  of  the  expansion  of  the  steam  occurs  in  the  moving 
parts.  At  the  present  time,  no  pure  reaction  turbines  are  used. 
The  so-called  reaction  turbine  in  use  expands  its  steam  both  in 
the  stationary  and  in  the  moving  parts.  It  therefore  employs  a 
mixture  of  the  impulse  principle  and  the  reaction  principle.  This 
mixed  type  is  shown  diagrammatically  in  Fig.  122.  The  steam 
enters  from  the  left  and  passes  through  a  set  of  stationary  blades 
in  which  there  occurs  some  drop  in  pressure.  In  passing  through 
the  next  set  of  blades,  which  are  moving,  a  further  drop  in  pres- 
sure occurs.  In  this  first  set  of  rotor  blades,  the  velocity  gen- 
erated in  the  stationary  blades,  and  also  that  produced  in  the 
rotor  blades  is  absorbed.  Hence  there  is  a  drop  in  pressure 


166 


ENGINES   AND   BOILERS 


throughout  the  whole  length  of  the  turbine.  As  the  velocity  is 
used  up  as  fast  as  it  is  produced,  no  very  high  steam  velocity 
exists  at  any  time  during  its  passage  through  the  turbine.  The 
graphs  show  the  drop  in  pressure  and  also  the  change  in  velocity 
as  the  steam  passes  through  the  turbine.  The  exact  ratio  of 
pressure-drop  or  velocity-change  is  not  shown  in  the  curves,  as 
the  scales  used  in  the  curves  on  the  small  cut  are  only  illustrative. 


Multi  3tage    Impure  Reaction 
FIG.  122 


It  has  been  shown  previously  (§  143)  that  the  velocity  of  the 
blades  of  a  pure  reaction  turbine  should  be  the  same  as  that  of 
the  steam  relative  to  the  blades,  in  order  to  obtain  the  highest 
efficiency.  This  means  that,  for  the  conditions  we  assumed  in 
§  147,  the  steam  velocity  to  be  used  up  per  stage  should  be  471. 
This  velocity  corresponds  to  a  kinetic  energy  per  pound  of  steam 
of 

=  3470  foot-pounds. 


This  is  equivalent  to  3470/778  =4.46  B.t.u.  per  pound  of  steam, 
which  is  one-fourth  of  the  value  that  was  obtained  for  the  im- 


STEAM   TURBINES 


167 


pulse  turbine.  For  a  heat  drop  of  175  B.t.u.  (165  to  15  pounds 
absolute),  it  will  require  175/4.46=39  stages.  Under  the  con- 
densing conditions  assumed  in  §  147,  323/4.46  =  72  stages 
would  be  necessary.  For  the  pressure-stage  impulse  type,  the 
values  were  10  and  18.  Since  the  reaction  turbine  under  discus- 
sion employs  a  combination  of  the  impulse  principle  and  the 
reaction  principle,  the  values  for  the  number  of  stages  may  be 
taken  as  a  mean  between  the  values  for  the  pure  impulse  type 
and  pure  reaction  type,  that  is  25  for  the  non-condensing  condi- 
tion, and  45  for  the  condensing  condition. 

The  preceding  computations  show  how  much  greater  length 
the  reaction  turbine  must  have  than  the  impulse  turbine.  The 
disadvantage  of  great  length  is  offset  to  some  extent  by  the  fact 
that  the  loss  due  to  friction  is  less  in  the  reaction  type,  since  the 
steam  velocities  are  low. 

151.  Summary.  —  To  recapitulate,  assuming  that  the  speed  is 
3000  r.  p.  m.,  and  that  the  diameter  of  the  rotor  is  3  feet,  the 
necessary  stages  for  each  of  the  various  types  is  shown  in  the 
following  table. 


Number  of  rows  of  buckets  or 
blades  on  the  rotor 

Heat-drop  of 
175  B.t.u. 

Heat-drop  of 
323  B.t.u. 

Velocity-stage  impulse 

3 
10 
4 

40 
25 

4 

18 
8 
72 
45 

Pressure-stage  impulse  .  . 

Mixed  pressure-  and  velocity-stage  impulse  . 
Pure  reaction  
Impure  reaction  

152.  Change  of  Area  of  Steam  Passage  Space.  —  It  is  neces- 
sary to  design  a  turbine  so  that  the  area  of  opening  for  the  passage 
of  steam  gives  the  proper  velocity  at  all  times.  As  the  pressure 
drops,  the  volume  of  steam  increases.  Hence  there  must  be  a 
very  much  larger  area  for  the  passage  of  steam  in  the  later  stages 
than  in  the  first  stages.  This  increase  is  accomplished  in  various 
ways.  If  the  diameter  of  all  the  rotor  wheels  is  kept  the  same, 
the  area  may  be  increased  by  having  only  partial  peripheral 
admission  of  the  steam  in  the  early  stages.  That  is,  the  nozzles 
or  stationary  vanes  may  extend  only  part  way  around  the  perim- 


168 


ENGINES   AND    BOILERS 


eter  of  the  stator.  In  Fig.  123  a  is  the  opening  for  the  passage 
of  steam  in  the  first  stage  of  the  turbine.  In  the  next  stage,  the 
opening  extends  farther  around  the  stator,  thus  giving  a  large 
area  for  the  passage  of  steam,  and  so  on  till,  in  the  later  stages, 
the  opening  extends  entirely  around  the  stator. 
The  area  may  be  increased  by  having  full 
peripheral  admission  in  all  the  stages,  but 
having  the  length  of  nozzles  and  blades  or 
buckets  increase  with  each  successive  stage. 
Figure  124  shows  this  scheme.  Or  if  full 
peripheral  admission  is  given  and  the  blades 
and  nozzles  are  kept  the  same  length,  the 
area  for  the  passage  of  steam  may  be  in- 
creased by  increasing  the  diameter  of  each  successive  stage,  as 
shown  in  Fig.  125.  ,Ih  practice,  combinations  of  these  methods 
of  increasing  the  area  for  the  passage  of  steam  are  used.  This 
increase  in  area  to  allow  for  the  increase  in  the  volume  of  steam 
passing,  must  not  be  confused  with  the  increase  in  area  in  a 


FIG.  123 


FIG.  124 


FIG.  125 


FIG.  126 


velocity-stage  turbine  which  is  necessary  to  allow  for  the  decrease 
in  steam  velocity  for  the  different  velocity  stages. 

153.  Leakage.  —  The  leakage  of  steam  through  an  opening 
depends  upon  the  difference  in  pressure  that  exists  on  the  two 
sides  of  the  opening,  and  upon  the  area  of  the  opening.  In  most 
steam  turbines,  there  are  necessarily  differences  in  pressure,  and 
openings  through  which  steam  may  escape.  With  the  rotor  of  a 
turbine  moving  at  a  high  speed,  it  is  not  possible  to  have  tight 
joints  at  all  places  between  the  stationary  and  the  moving  parts, 
for  then  the  friction  between  these  parts  would  create  an  even 


STEAM    TURBINES  169 

greater  loss.  In  design  and  construction  the  clearances  are  kept 
as  small  as  is  consistent  with  economy  and  safety,  but  even  then 
some  leakage  is  sure  to  occur. 

In  Fig.  126,  a  difference  in  pressure  exists  between  the  two 
sides  of  the  stator  at  c  and  d;  hence  there  will  be  a  leakage  of 
steam  at  a.  If  there  is  a  difference  in  pressure  between  d  and  e, 
a  leakage  occurs  at  6,  and  so  on  for  the  various  stages.  In  the 
velocity-stage  impulse  turbine  there  is  no  difference  in  pressure 
after  the  steam  leaves  the  nozzles;  hence  there  is  no  tendency 
to  leak  at  either  a  or  b.  In  the  pressure-stage  impulse  type,  there 
is  a  difference  in  pressure  between  c  and  d,  and  therefore  leakage 
at  a;  but  there  is  none  between  d  and  e. 

In  the  impure  reaction  type  there  is  a  difference  in  pressure 
between  c  and  d,  and  also  between  d  and  e;  hence  there  is  leakage  at 
both  a  and  b.  The  leakage  at  a  can  be  kept  at  a  minimum  by 
making  the  opening  there  as  small  as  possible.  This  opening  at  a 
depends  upon  the  closeness  of  the  fit  and  upon  the  diameter  of 
the  shaft.  Carbon  or  other  form  of  packing  is  sometimes  used 
here  to  give  a  close  fit.  In  most  reaction  turbines,  however,  the 
shaft  is  very  large,  in  fact  it  is  even  a  drum,  and  in  that  case  the 
area  of  the  opening  is  quite  large.  Packing  cannot  well  be  used 
with  large  diameters. 

In  the  reaction  turbine  the  difference  in  pressure  is  not  so 
great  between  c  and  e  as  in  the  impulse  type.  In  the  reaction 
turbine  a  leakage  occurs  at  6;  hence  the  clearance  there  is  kept  as 
small  as  is  consistent  with  safety  and  economy.  A  certain  amount 
of  spillage  occurs  at  b  because  the  rotor  throws  some  of  the  steam 
out  by  centrifugal  force  and  it  escapes  through  g,  even  if  there 
is  no  difference  in  pressure  between  d  and  e. 

154.  Loss  Due  to  Running  at  Partial  Capacity.  —  Every  tur- 
bine is  designed  to  expand  the  steam  from  a  certain  initial  pres- 
sure to  a  certain  back  pressure.  This  does  not  mean  that  the 
turbine  will  run  only  under  the  pressure  conditions  assumed  in 
its  design,  but  it  does  mean  that  the  efficiency  will  be  low  if 
there  is  any  great  difference  from  the  assumed  conditions.  For 
instance,  if  a  turbine  is  designed  to  run  non-condensing,  and  is 
operated  condensing,  it  means  that  the  efficiency  will  be  consid- 
erably less  than  it  would  be  if  the  turbine  had  been  designed  to 
run  condensing.  The  reason  for  this  is  easy  to  understand. 


170  ENGINES   AND    BOILERS 

Every  nozzle  and  each  steam  passage  is  designed  to  carry  a  cer- 
tain volume  of  steam  with  a  certain  velocity.  We  have  shown 
that  velocity  is  caused  by  drop  in  pressure,  so  that  if  the  pressure- 
drops  are  not  as  designed,  the  velocities  are  not  such  that  the 
efficiency  will  be  a  maximum. 

When  a  turbine  is  designed  for  a  certain  load,  and  that  load 
is  greatly  increased  or  decreased,  it  is  evident  that  the  efficiency 
will  be  decreased.  In  the  design  of  the  governor  the  aim  is  to 
make  this  drop  in  efficiency  as  small  as  possible,  while  at  the  same 
time  maintaining  uniform  speed  for  all  loads.  It  is  also  essential 
that  the  governor  be  reliable  and  therefore  not  too  complicated. 
The  simplest  form  of  governor  is  a  throttling  governor,  and  many 
of  the  smaller  turbines  are  equipped  with  that  kind.  However,  with 
light  loads,  the  throttling  governor  causes  a  large  loss  in  efficiency 
because  the  range  in  pressure  between  admission  and  exhaust 
is  so  largely  decreased. 

On  the  larger  machines  some  other  means  than  throttling  us- 
ually is  used.  If  the  machine  is  of  the  impulse  type,  it  has  sta- 
tionary nozzles,  and  the  governor  can  be  arranged  to  open  the 
proper  number  to  take  care  of  the  load.  If  the  machine  had  a 
single  stage,  or  if  a  proportional  number  of  nozzles  could  be  kept 
open  in  the  later  stages  of  a  multi-stage  turbine,  this  would  seem 
to  be  an  ideal  arrangement.  But  the  large  machines  never  are 
made  single-stage  and  a  governor  to  control  all  the  nozzles  in 
all  the  stages  would  be  very  complicated. 

In  practice,  as  in  the  ordinary  Curtis  turbine,  the  governor 
controls  the  nozzles  of  the  first  stage  only.  At  light  loads  only 
a  few  nozzles  of  the  first  stage  are  open,  while  at  maximum  load 
all  the  first-stage  nozzles  are  open.  This  arrangement  assures 
good  economy  in  the  first  stage  at  light  loads,  but  as  all  the  later- 
stage  nozzles  are  open,  there  will  be  a  loss  there. 

In  the  single-stage  DeLaval  turbine,  a  throttling  governor  is 
used,  but  an  effort  is  made  to  maintain  the  best  economy  by 
having  some  of  the  nozzles  controlled  by  hand-operated  valves. 
By  this  method,  most  of  the  nozzles  can  be  shut  off  at  small 
load  and  opened  up  by  hand  for  heavy  load.  The  governor  is 
incapable  of  controlling  the  load  entirely  if  manual  control  is 
attempted. 

In  a  reaction  type  of  machine  in  which  there  are  no  stationary 
nozzles,  the  preceding  method  of  governor-control  cannot  be  used. 


STEAM   TURBINES  171 

The  Westinghouse  Company,  on  their  Parsons  type  of  turbine, 
has  attempted  to  secure  good  efficiency  at  light  loads  by  having 
the  governor  admit  the  steam  in  puffs,  which  is  the  plan  intro- 
duced by  Parsons.  That  is,  the  steam  is  admitted  at  full  pres- 
sure for  a  short  time  and  then  entirely  cut  off.  The  interval 
between  puffs  is  practically  constant,  but  the  length  of  time  the 
full  steam  pressure  is  on  during  the  puff  is  controlled  by  the 
governor.  In  the  later  stages,  the  effect  of  this  kind  of  governor 
approximates  that  of  a  plain  throttling  governor.  In  other  makes 
of  reaction  turbine  a  throttling  governor  is  commonly  used. 

Large  overloads  are  often  carried  in  the  various  types  of  tur- 
bines by  turning  full  steam  pressure  into  a  later  stage.  In  this 
way  the  machine  is  able  to  carry  a  large  excess  of  load  at  a  re- 
duced efficiency.  If  an  electric  generator  is  attached  to  the  tur- 
bine, care  must  be  exercised  that  it  is  not  overloaded  long 
enough  to  get  too  hot.  This  method  is  used  only  in  emergencies, 
and  then  only  for  short  periods  of  time. 

155.  Summary  of  Losses  in  the  Steam  Turbine.  —  The  losses 
that  occur  in  a  turbine  have  been  mentioned  in  §§  140,  147, 
153,  154,  and  elsewhere.  We  shall  now  make  a  summary  of  the 
more  serious  causes  of  loss. 

FRICTION  LOSSES.    Losses  due  to  friction  occur  as  follows : 

(1)  Between  the  shaft  and  the  bearings,  and  in  the  packing 
rings  where  the  turbine  is  made  steam-tight.     With  proper  design 
and  construction  this  loss  is  quite  small. 

(2)  Between  the  steam  and  parts  of  the  turbine.     The  steam 
friction  varies  directly  as  the  steam  pressure,  and  as  the  square 
of  the  velocity  between  the  steam  and  the  parts.     It  also  varies 
as  the  amount  of  surface  in  contact  with  the  steam.     Hence  the 
amount  of  surface  in  contact  with  the  steam  should  be  kept  as 
small  as  possible  and  the  velocity  should  be  kept  as  low  as  possible. 
The  steam-friction  loss  in  the  first  stages  may  be  partially  reclaimed 
in  the  later  stages  since  the  heat  generated  tends  to  raise  the  tem- 
perature of  the  steam.     The  steam  friction  occurs  as  the  steam 
flows  through  the  nozzles  and  blades. 

(3)  There  is  also  a  friction  loss  between  the  rotor  discs  and  the 
steam  surrounding  them,  at  d,  e,  /,  and  g  in  Fig.  126.     This  loss 
is  called  windage;  it  may  be  reduced  by  having  the  rotor  discs 
smooth  and  polished. 


172  ENGINES   AND   BOILERS 

LEAKAGE  LOSSES.  Leakage  occurs  wherever  a  difference  in  pres- 
sure exists  on  the  two  sides  of  an  opening.  It  is  thus  seen  that 
there  may  be  a  leakage  out  of  the  casing  or  through  the  joints 
between  the  pressure  stages,  or  around  the  balance  pistons  of  the 
reaction  turbine  (§  160). 

Another  loss  occurs  under  condensing  conditions  because  air 
leaks  into  the  low-pressure  parts  of  the  turbine.  This  air  tends  to 
lower  the  vacuum  in  the  condenser,  and  may  thereby  cause  a  loss. 

DISTURBANCE  OF  FLOW.  A  breaking  up  of  the  jet  of  steam,  and 
the  consequent  formation  of  eddies,  causes  a  considerable  loss  in 
some  types  of  turbines.  This  is  one  of  the  causes  of  the  low 
efficiency  of  the  velocity-stage  impulse  turbine  (§  147). 

LACK  OF  PROPER  VELOCITY.  It  has  been  shown  that  the  proper 
relation  must  exist  between  the  velocity  of  the  buckets  and  that 
of  the  jet  to  obtain  the  highest  efficiency  (§  143).  If  the  velocity 
of  the  blades  or  buckets  is  not  correct,  there  will  be  a  loss.  Throt- 
tling of  steam  by  the  governor  will  cause  a  loss.  (See  §  154.) 

EXIT  VELOCITY.  The  turbine  derives  its  energy  by  absorbing 
the  steam  velocity.  A  high  velocity  of  steam  in  the  exhaust  en- 
tails a  decided  loss.  The  turbine  should  be  designed  and  operated 
to  extract  nearly  all  the  steam  velocity,  and  to  leave  only  enough 
in  the  exhaust  to  cause  the  steam  to  flow  to  the  condenser. 

156.  Common  Commercial  Types.  —  A  great  many  makes  of 
turbines  have  been  used  in  the  past.     Some  of  these  were  not 
economical  and  have  been  replaced.     Others  have  ceased  to  exist 
for  other  reasons.     At  the  present  time  there  are  a  great  many 
different  forms  in  use,  but  space  will  not  permit  us  to  consider 
all  of  them.     Some  of  the  more  common  forms  will  be  explained. 

157.  DeLaval  Single-stage  Steam  Turbine.  —  The   DeLaval 
single-stage  turbine  is  the  oldest  of  the  types  used  at  present. 
It  dates  back  to  the  years  1880-1890.    DE  LAVAL,  the  inventor  of 
the  cream  separator,  sought  to  drive  that  device  by  means  of  a 
direct-connected  turbine.     In  his  experiments  he  developed  the 
type  that  bears  his  name.     Some  minor  changes  have  been  made, 
but  its  essential  features  remain  the  same.     The  DeLaval  tur- 
bine used  in  America  is  manufactured  by  an  American  company 
that   originally   produced    their   machines   under   the   DeLaval 
patents.     The  same  company  now  makes  a  multi-stage  machine, 
but  it  is  not  to  be  confused  with  the  single-stage  type. 


STEAM   TURBINES 


173 


The  essential  features  of  the  DeLaval  turbine  are  as  follows: 

(1)  The  expanding  nozzle  in  which  all  the  pressure-drop  occurs. 

(2)  The  rotor  wheel  which  carries  a  single  row  of  buckets. 

(3)  A  slender,  flexible  shaft  that  carries  the  wheel  and  transmits 
the  power  to  the  gears. 

(4)  A  set  of  reduction  gears  which  lowers  the  speed  so  that  it 
is  usable. 

One  style  of  the  DeLaval  turbine  is  shown  in  Fig.  127.     The 


FIG.  127 

bucket  wheel  R  is  mounted  on  a  flexible  pinion  shaft  which  is 
supported  by  the  bearings  B,  B'.  This  small  shaft  also  carries 
the  small  pinion  P  which  is  supported  by  the  bearings  &',  6'. 
The  pinion  meshes  with  the  two  large  helical  gears  shown  in 
the  figure.  The  shafts  that  carry  the  large  gears  are  supported 
by  the  bearings  b,  b.  The  shafts  of  the  electric  generators  or  pumps 
to  be  driven  by  the  turbine  are  coupled  to  the  large-gear  shafts. 
The  governor  also  is  driven 


from  one  of  these  shafts. 

The  expanding  nozzle  is 
shown  in  Fig.  128.  The  steam 
enters  from  the  left  and  the 


FIG.  128 


area  of  the  opening  gradually  increases  to  a  size  that  produces  the 
proper  pressure  and  velocity.  The  amount  of  flare  given  to 
the  nozzle  is  governed  by  the  drop  in  pressure  that  is  desired. 


174  ENGINES   AND    BOILERS 

The  nozzle  used  if  the  turbine  exhausts  to  the  atmosphere  is 
called  a  non-condensing  nozzle.  If  the  turbine  exhausts  into  a 
condenser,  the  nozzle  is  called  a  condensing  nozzle.  The  flare  in  the 
non-condensing  nozzle  is  less  than  that  in  the  condensing  nozzle. 

One  or  more  of  the  nozzles  are  open  at  all  times;  the  rest  are 
opened  or  closed  by  hand,  depending  upon  the  amount  of  the 
load.  The  governor  is  of  the  centrifugal  type,  and  governs  by 
throttling  the  steam.  The  nozzles  are  placed  in  the  casing  par- 
tition at  N  (Fig.  127).  The  steam  chest  is  at  $;  the  steam  passes 
from  it  through  the  nozzles,  then  through  the  row  of  buckets  on 
the  rotor,  and  out  into  the  exhaust  space  E.  From  E,  the  steam 
is  led  to  the  condenser  or  to  the  atmosphere. 

We  saw  in  §  145  that  the  bucket  velocity  of  a  single-stage  tur- 
bine must  be  very  high  in  order  to  extract  a  reasonable  part  of 
the  kinetic  energy  of  the  steam.  The  high  bucket-velocity  causes 
large  stresses  in  the  rotor.  The  rotor  disc  therefore  is  made  of  high 
quality  alloy  steel  and  is  very  carefully  designed  and  constructed. 
A  sufficient  factor  of  safety  exists  at  the  rated  speed,  but  any 
large  increase  above  this  speed  will  cause  failure  in  the  wheel. 
It  is  customary  to  make  the  rotor  weakest  at  a  point  just  inside 
the  rim  where  the  buckets  are  attached,  so  that,  if  the  breaking 
speed  is  reached,  only  the  rim  of  the  rotor  will  tear  loose.  When 
that  happens,  the  speed  will  die  down  of  itself  because  the  buckets 
are  gone.  If  the  shaft  becomes  sprung,  the  safety  bearings  around 
the  hub  keep  the  main  part  of  the  rotor  in  place,  so  that  no  great 
damage  results.  The  steel  casing  around  the  rotor  is  strong 
enough  to  keep  any  small  fragments  from  breaking  through;  but  if 
the  whole  rotor  should  break  in  two,  it  is  likely 
that  considerable  damage  would  be  done. 

The  buckets  are  drop-forged.  They  are  at- 
tached to  the  wheel  as  shown  in  Fig.  129. 
Slots  are  machined  into  the  perimeter  of  the 
rotor  and  the  buckets  are  forced  into  the  slots 
from  the  side.  Should  an  individual  bucket 
become  damaged,  it  may  be  removed  and  an- 
FIG.  129  other  put  in  its  place. 

At  high  speed,  there  is  a  tendency  for  the 
rotor  to  vibrate  because  it  is  impossible  to  get  the  center  of  gravity 
exactly  in  the  center  of  rotation.  There  is  a  critical  speed  at  which 
the  vibration  is  a  maximum.  At  a  speed  above  this  critical  value, 


JM*!M  of  a-t 


STEAM   TURBINES  175 

the  shaft  or  the  bearings  yield  slightly,  and  the  center  of  gravity 
of  the  rotor  comes  into  the  axis  of  rotation.  The  rotor  then  runs 
smoothly  again.  For  smooth  running,  the  speed  must  be  either 
very  much  less,  or  considerably  greater,  than  the  critical  value. 
Making  the  shaft  small  tends  to  reduce  the  critical  speed.  The 
shaft  of  this  turbine  is  made  very  small  so  that  it  may  give 
easily,  and  so  run  smoothly  at  its  desired  speed,  which  is  normally 
above  the  critical  speed.  While  the  diameter  of  the  shaft  is  small,  it 
is  ample  to  carry  the  power  at  the  high  velocity  at  which  it  runs. 

The  rotor  speed  in  the  smallest  sizes  is  as  high  as  30,000  revo- 
lutions per  minute.  For  a  300-horsepower  turbine,  the  speed  is 
10,000  revolutions  per  minute.  It  will  be  seen  that  it  is  imprac- 
ticable to  run  an  electric  generator  or  a  pump  at  such  speed, 
and  to  utilize  the  power  developed,  a  speed  reduction  must  be 
used.  Helical  gears  are  used  for  this  reduction  in  order  to  insure 
smooth,  quiet  running.  In  some  models,  one  large  gear  is  used; 
in  others  there  are  two,  as  shown  in  Fig.  127.  The  latter  neces- 
sitates two  generators  or  pumps.  When  the  gears  are  new,  the 
loss  of  power  in  the  reduction  is  small.  This  loss  increases  as 
the  wear  increases.  Reduction  gears  are  now  used  in  some  other 
turbines,  though  they  were  used  originally  only  in  the  DeLaval 
turbine. 

Aside  from  its  use  in  the  cream  separator,  the  single-stage  De- 
Laval  turbine  has  been  used  in  driving  electric  generators,  cen- 
trifugal pumps  and  blowers.  It  is  not  used  in  sizes  above  500 
horsepower.  The  size  is  limited  because  the  buckets  would  have 
to  be  unreasonably  long,  or  else  the  diameter  of  the  rotor  would 
have  to  be  too  large,  in  order  to  get  enough  nozzles  to  play  on 
the  buckets, 

158.  The  Multi-pressure-stage  Impulse  Turbine.  —  We  have 
seen  that  there  is  a  practical  limit  to  the  size  of  single-stage  tur- 
bines. Because  it  gives  the  designer  greater  liberty  of  choice 
of  speeds,  velocities,  and  bucket  lengths,  the  multi-stage  impulse 
turbine  is  quite  common  in  small  and  medium  sizes.  Moreover, 
the  efficiency  of  this  type  is  comparatively  high  in  these  sizes. 
For  these  reasons,  it  is  quite  common  in  sizes  up  to  5000  horse- 
power. 

Since  it  has  a  number  of  pressure  stages  or  cells,  this  kind  of 
turbine  is  sometimes  called  the  multi- cellular  type.  It  is  also 


176 


ENGINES   AND   BOILERS 


called  the  Rateau  type,  because  RATEAU  was  the  first  to  develop 
it,  but  there  is  no  essential  difference  between  the  Eateau  tur- 
bine and  the  numerous  other  multi-cellular  turbines. 

Figure  130  shows  an  Economy  turbine,  made  by  the  Kerr  Tur- 
bine Company.  As  seen  in  the  figure,  there  is  a  rotor  shaft  which 
carries  a  number  of  rotor  wheels  or  discs.  Each  wheel  runs  in  a 
pressure  cell.  The  cells  are  separated  by  the  heavy  diaphragms 
shown  at  a.  The  joint  between  the  diaphragm  and  the  shaft  is 
kept  as  nearly  steam-tight  as  possible  by  making  as  close  a  fit 
as  is  practicable  under  the  circumstances.  The  buckets  are  fastened 


FIG.  130 

on  the  rim  of  the  wheels  in  much  the  same  manner  as  in  the  single- 
stage  DeLaval  turbine.  The  shaft  is  supported  by  the  bearings 
B,  B.  Nozzles  are  located  in  the  cell  partitions  as  at  N.  The 
number  of  nozzles  increases  from  the  first  to  the  last  stage,  in 
order  to  allow  for  the  increased  volume  of  the  steam  as  the  pressure 
drops. 

The  steam  enters  the  steam  chest  S,  and  passes  through  the 
nozzle  Nt  where  it  is  partially  expanded.  Leaving  the  nozzle,  it 
passes  through  the  first  row  of  buckets  6,  on  the  rotor,  then 
through  the  second  set  of  nozzles,  and  so  on  till  it  arrives  at  the 
exhaust  E.  Leakage  of  steam  from  the  high-pressure  end  of  the 
turbine,  and  of  air  into  the  exhaust,  is  prevented  by  the  stuffing- 
boxes  G,  G. 

The  casing  is  lagged  with  a  non-conducting  material  to  prevent 
the  loss  of  heat  by  radiation.  The  bearings  are  equipped  with 


STEAM   TURBINES  177 

ring  oilers.  As  in  other  impulse  turbines,  there  is  no  difference 
between  the  pressure  on  the  two  sides  of  the  rotor.  Hence  there 
is  no  tendency  to  leak  steam  over  the  ends  of  the  buckets,  and  it 
is  unnecessary  to  make  the  clearances  between  the  buckets  and 
the  stationary  elements  small.  This  obviates  the  necessity  of 
careful  adjustments,  and  the  danger  of  contact  due  to  any  un- 
equal expansion  of  the  parts.  The  diaphragms  which  compose 
the  cell  walls  are  made  rigid  so  that  they  will  not  spring  as  a 
consequence  of  the  differences  in  pressure  on  their  opposite  sides. 

The  turbine  is  supplied  with  a  throttling  governor  of  the  cen- 
trifugal type,  driven  from  a  worm  on  the  main  shaft.  In  the 
smaller  sizes  it  is  direct-acting,  that  is,  the  position  of  the  gov- 
ernor weights  directly  controls  the  opening  of  the  throttle  valve. 
In  the  larger  sizes,  it  has  been  found  desirable  to  have  a  relay 
arrangement  by  which  the  throttle  valve  is  thrown  by  steam  or 
by  hydraulic  pressure,  which  in  turn  is  controlled  by  the  position 
of  the  governor  weights.  This  arrangement  gives  better  speed 
regulation  and  does  not  require  so  large  a  governor.  Consider- 
able force  is  required  to  operate  the  large  throttle  valves. 

Machines  of  this  type  made  by  other  companies  closely  re- 
semble the  one  just  mentioned  in  essential  features.  Designers 
often  increase  the  length  of  the  blades  or  buckets  with  each  succes- 
sive stage,  thus  helping  to  allow  for  the  increase  in  volume.  The 
diameter  of  rotor  wheels  is  sometimes  made  larger  in  the  later 
stages  for  the  same  purpose.  Various  devices  are  used  to  keep  the 
leakage  from  stage  to  stage  at  a  minimum. 

Carbon  packing  rings  are  sometimes  used.  In  other  makes, 
labyrinth  joints  are  used.  No  large  mechanical  pressure  is  allow- 
able at  the  contact  points  between  stages.  The  result  is  that 
there  is  often  a  considerable  loss  from  leakage  of  steam  around 
the  wheels,  sometimes  as  high  as  15  to  20  per  cent. 

Thrust  bearings  are  also  used  to  prevent  end-play  of  the  shaft, 
which  might  cause  the  buckets  to  rub  and  to  become  injured. 

The  practice  of  cutting  holes  in  the  rotor  disc  to  insure  an  equal- 
ization of  pressure  on  the  two  sides  of  the  rotor  is  common.  The 
shaft  should  be  made  large  enough  so  that  the  speed  at  which  it 
is  to  run  is  well  below  the  critical  speed,  and  then  excessive  vibra- 
tion will  not  occur.  It  is  necessary  to  use  good  workmanship  and 
make  the  rotor  well  balanced;  otherwise,  severe  strains  would  be 
induced  at  the  high  speeds  at  which  the  turbine  is  run. 


178  ENGINES   AND   BOILERS 

It  is  not  our  purpose  to  describe  the  details  of  construction  of 
all  the  various  makes  of  turbines.  It  must  be  remembered  that 
each  make  has  its  own  minor  variations  in  construction,  but  the 
preceding  description  holds  in  general  for  all  makes  of  this  type  of 
machine. 

159.  The  Curtis  Steam  Turbine.  —  The  original  patents  for  the 
Curtis  turbine  were  issued  about  1895,  and  the  General  Electric 
Company  started  production  of  this  machine  shortly  afterwards. 
For  several  years  it  was  built  in  this  country  only  by  them.  More 
recently,  however,  the  Curtis  principle  has  been  used  largely  in. 
the  high-pressure  stages  of  some  other  machines.  At  first  the 
General  Electric  Company  built  the  larger  turbines  with  the  axis 
of  the  rotor  vertical.  The  generator,  which  was  direct-connected 
to  the  rotor  shaft,  was  placed  on  top  of  the  turbine,  and  the  con- 
denser was  placed  directly  beneath  it.  This  was  claimed  to  be  an 
ideal  arrangement,  and  in  some  respects  it  was  excellent,  but  all  the 
weight  of  the  rotating  parts  of  both  the  generator  and  the  turbine 
had  to  be  supported  by  a  step-bearing.  With  every  precaution, 
and  the  best  of  design,  this  step-bearing  would  sometimes  fail,  and 
this  failure  often  caused  the  buckets  to  strip.  For  very  large, 
fast-speed  turbines,  it  was  difficult  to  secure  sufficient  mechanical 
rigidity  for  the  bearing  supports  in  vertical  machines.  The  makers 
therefore  have  come  to  prefer  the  horizontal  type.  A  great  many 
of  the  vertical  turbines  are  still  in  successful  operation,  however. 
The  following  description  is  taken  from  General  Electric  Co. 
Bulletin  No.  4883. 

Figure  131  shows  diagrammatically  the  progress  of  the  steam 
in  a  Curtis  turbine.  Entering  at  A  from  the  steam  pipe,  it 
passes  into  the  steam  chest  B,  and  then  through  one  or  more 
open  valves  to  the  bowls  C.  The  number  of  valves  open  de- 
pends upon  the  load,  and  their  action  is  controlled  by  the  gov- 
ernor. From  the  bowls  C,  the  steam  expands  through  diverg- 
ing nozzles  D,  entering  the  first  row  of  revolving  buckets  of 
the  first  stage  at  E,  then  passing  through  the  stationary  buckets 
F,  which  reverse  its  direction  and  redirect  it  against  the  second 
revolving  row  G. 

This  constitutes  the  performance  of  the  steam  in  one  stage, 
or  pressure  chamber.  Having  entered  the  first  row  of  buckets 
at  E  with  relatively  high  velocity,  it  leaves  the  last  row  G 


STEAM   TURBINES 


179 


with  a  relatively  low  velocity,  its  energy  between  the  limits 
of  inlet  and  discharge  pressure  having  been  extracted  in  passing 
from  C  to  H.  It  has,  however,  a  large  amount  of  unexpended 
energy,  since  the  expansion  from  C  to  E  has  covered  only  a  part 
of  the  available  pressure-range.  The  expansion  process  is,  there- 
fore, repeated  in  a  second  stage. 

The  steam,  having  left  the  buckets  G,  and  having  had  its  ve- 
locity greatly  reduced,  reaches  a  second  series  of  bowls  H, 
opening  upon  a  second  series  of  nozzles  /.  Through  these  the 


FIG.  131 

steam  expands  again  from  the  first-stage  pressure  to  some  lower 
pressure,  again  acquiring  relatively  high  velocity  in  its  expan- 
sion through  these  nozzles,  leaving  them  at  J  and  impinging 
upon  and  passing  through  the  moving  and  stationary  buckets 
K,  L  and  M,  precisely  as  in  the  first  stage.  Again  the  velocity 
acquired  in  the  nozzle  is  expended  in  passing  through  the  mov- 
ing and  stationary  buckets,  and  the  steam  leaves  the  second  row 
M  with  relatively  low  velocity. 

This  process  is  continued  in  most  large  turbines  through 
several  stages.     Curtis  machines  as  now  constructed  have  a 


180  ENGINES   AND    BOILERS 

single  stage  in  the  very  small  sizes,  up  to  six  or  seven  stages  in 
the  larger  ones. 

Again  referring  to  Fig.  131,  it  should  be  noticed  particularly 
that  the  pressure  of  the  steam  has  not  changed  in  its  passage 
from  E  to  H,  that  is  the  pressure  is  practically  constant  at  all 
points  in  the  stage.  This  fact  leads  to  one  of  the  principal 
structural  advantages  of  the  Curtis  type.  .  .  .  For,  since  the 
pressure  is  uniform  at  all  points,  there  is  no  tendency  for  the 
steam  to  pass  elsewhere  than  where  directed  by  the  nozzles, 
i.e.  through  the  buckets.  Hence  there  is  no  necessity  of  main- 
taining a  close  clearance  between  the  ends  of  the  revolving 
buckets  and  the  turbine  casing.  In  practice  free  clearance  is 
provided  from  one  to  two  inches.  The  reaction  type,  to  se- 
cure high  economy,  must  be  provided  with  a  minimum  clearance 
at  this  point. 

Also  since  the  pressure  on  both  sides  of  each  wheel,  i.e.  at 
E  and  H,  Fig.  131,  is  the  same,  the  wheel  is  in  perfect  equilib- 
rium, there  being  no  tendency  for  the  steam  to  force  the  wheel 
in  an  axial  direction.  As  this  is  true  of  each  wheel,  the  entire 
rotor  is  in  equilibrium,  and  there  is  practically  no  unbalanced 
thrust. 

As  the  steam  expands  from  stage  to  stage,  its  volume  rapidly 
increases,  and  a  greater  area  of  steam  passage  must  be  pro- 
vided. This  is  accomplished  in  two  ways.  First,  by  increas- 
ing the  height  of  the  buckets  and  second,  by  increasing  the 
number  and  area  of  nozzles  from  stage  to  stage.  Referring  to 
Fig.  131,  it  should  be  noted  that  the  primary  admission  noz- 
zles D  actually  extend  around  a  small  portion  of  the  first  stage 
periphery;  therefore  only  those  buckets  adjacent  to  the  nozzles 
at  any  instant  are  carrying  active  steam.  This  applies  equally 
to  the  stationary  row  and  the  second  revolving  row;  in  fact, 
the  stationary  or  intermediate  buckets,  as  built,  extend  over  a 
small  arc  not  much  larger  than  the  nozzle  arc.  In  the  second 
stage,  however,  the  nozzle  arc  becomes  longer  and  wider,  thus 
permitting  the  flow  of  steam  through  a  greater  number  of  re- 
volving buckets  and  necessitating  a  longer  arc  of  stationary 
buckets.  Finally,  when  the  low-pressure  stages  are  reached, 
the  nozzles  and  stationary  buckets  extend  all  the  way  around 
the  circumference. 

As  previously  mentioned,  greater  area  for  the  steam  flow  is 


STEAM   TURBINES  181 

also  provided  by  increasing  the  bucket  lengths.     For  example, 

the  first-stage,  or  high-pressure,  buckets  are  generally  less  than 

an  inch  long,  while  those  in  the  low-pressure  stages  may  be 

eight  or  ten  inches  in  length. 

It  will  be  noticed  that  the  length  of  the  second  row  of  moving 
buckets  in  each  stage,  G  and  M,  is  greater  than  that  of  the  first 
row.  This  is  made  so,  not  to  allow  for  any  expansion  of  the 
steam,  but  to  provide  for  a  decrease  or  velocity  of  the  same  vol- 
ume of  steam. 

Customarily  the  buckets  of  the  Curtis  turbine  are  dovetailed 
into  the  rim  of  the  rotor  wheel,  somewhat  as  shown  in  Fig.  131. 
At  intervals  the  dovetail  channel  in  the  rim  of  the  rotor  is  open 
for  the  insertion  of  buckets.  These  openings  are  afterwards  filled 
with  a  spacing  blank,  and  closed  up.  After  the  buckets  are  as- 
sembled a  shroud  ring  is  riveted  to  their  outer  ends.  The  func- 
tion of  this  ring  is  partly  to  stiffen  the  complete  row  and  to  reduce 
vibration,  but  more  especially  to  assist  in  retaining  the  steam 
flow  in  the  bucket  space.  Centrifugal  force  tends  to  throw  the 
steam  out  to  the  end  of  the  bucket. 

The  governor  is  of  the  centrifugal  type  and  controls  the  steam 
supply  by  opening  and  closing  some  of  the  nozzles  of  the  first 
stage.  Those  nozzles  that  are  open,  are  wide  open,  and  those 
that  are  closed,  are  tight  shut.  This  scheme  is  positive  and 
reliable.  It  gives  close  speed  regulation,  and  high  efficiency  at 
light  loads.  In  addition  to  the  governor,  the  machine  is  equipped 
with  an  emergency  stop,  whose  function  is  to  prevent  excessive 
speeds,  should  the  governor  fail  for  any  reason.  It  consists  of 
an  unequally  weighted  ring  attached  to,  and  revolving  with,  the 
shaft.  At  any  speed  up  to  the  normal  speed,  the  weights  are  held 
concentric  with  the  shaft  by  springs,  but  at  excessive  speeds  the 
force  of  the  springs  is  overcome.  Then  the  ring  revolves  eccen- 
trically, and  trips  the  valve  mechanism,  causing  the  main  throttle 
valve  to  close  instantly,  thereby  shutting  off  the  steam  supply. 

Figure  132  shows  a  marine  Curtis  turbine.  In  marine  service, 
the  speed  must  be  very  much  less  than  is  common  in  land  prac- 
tice, if  the  rotor  is  coupled  directly  to  the  propeller  shaft.  To 
get  this  low  speed,  it  is  necessary  to  use  very  large  diameters  or 
a  great  number  of  stages.  Usually  both,  schemes  are  combined. 
As  has  been  stated  previously,  the  efficiency  after  the  second 
stage  of  a  velocity-staged  impulse  turbine  is  low,  but  in  order  to 


182 


STEAM   TURBINES  183 

reduce  the  speed  properly,  it  may  be  desirable  to  use  more  than 
two  velocity  stages,  even  if  the  efficiency  is  reduced.  In  land 
practice  this  concession  is  not  often  made.  In  the  forward  tur- 
bine shown  in  Fig.  132,  there  are  four  velocity  stages  in  the  first 
pressure  stage,  three  velocity  stages  in  each  pressure  stage  from 
the  second  to  the  sixth,  and  two  velocity  stages  in  each  pressure 
stage  from  the  seventh  to  the  fourteenth.  In  the  reverse  tur- 
bine, there  are  four  velocity  stages  in  each  of  the  pressure  stages. 

It  is  customary  in  marine  turbines  to  have  the  reverse  turbine 
mounted  on  the  same  shaft  with  the  forward  or  ahead  turbine. 
It  is  put  on  the  exhaust  end  of  the  shaft  so  that  when  it  runs 
idle,  the  rotor  is  in  a  high  vacuum  and  therefore  offers  as  little 
resistance  as  possible  to  rotation.  Since  the  friction  loss  varies 
as  the  steam  pressure,  the  loss  is  not  great  for  low  pressures. 

The  efficiency  during  reverse  operation  is  quite  poor  because 
there  are  only  two  pressure  stages  in  the  reverse  turbine.  The 
reverse  is  in  use  only  briefly  and  rather  unfrequently.  Hence  the 
the  low  efficiency  of  the  reverse  turbine  is  a  negligible  factor. 

After  a  study  of  the  previous  types,  Fig.  132  should  be  largely 
self-explanatory,  hence  no  detailed  description  need  be  given. 
It  is  to  be  noted  that  the  buckets  of  the  seventh  to  fourteenth 
stages  are  carried  by  a  drum.  As  the  pressure  on  the  right  end 
of  this  drum  is  greater  than  that  on  the  left  end,  it  is  seen  that 
there  will  be  an  end-thrust  of  the  shaft.  This  thrust  is  taken 
up  by  the  thrust  bearing  T.  The  thrust  bearing  T  is  for  the 
turbine  only,  and  not  for  the  propeller.  On  every  propeller  shaft 
there  is  a  thrust  bearing  to  take  up  the  thrust  of  the  propeller. 
The  bearings  are  provided  with  a  water  jacket  to  prevent  heating. 
In  all  large  turbines,  the  bearings  are  cooled  either  by  means  of 
water  or  oil.  When  oil  is  used,  it  is  often  cooled  by  a  device 
similar  to  the  surface  condenser. 

160.  The  Parsons  Steam  Turbine.  —  The  Parsons  turbine  is 
not  only  one  of  the  oldest,  but  also  one  of  the  most  common  types. 
It  is  usually  made  in  medium  and  large  sizes.  In  small  sizes  the 
Parsons  turbine  is  expensive,  and  is  not  very  efficient.  All  of  the 
previous  types  have  operated  on  the  impulse  principle,  but  this 
one  uses  a  mixture  of  impulse  and  reaction.  It  is  ordinarily 
called  a  reaction  turbine.  There  are  no  distinct  nozzles,  as  in 
the  impulse  turbine.  Instead,  there  are  alternate  fixed  and  mov- 


184 


STEAM   TURBINES  185 

ing  blades,  and  expansion  occurs  in  both.     It  is  made  by  the 
Westinghouse  Company  and  by  the  Allis-Ch aimers  Company. 

Figure  133  shows  one  style  of  Parsons  Turbine.  Instead  of 
rotor  wheels  as  in  the  previous  types,  the  shaft  carries  a  drum 
on  which  the  moving  blades  are  mounted.  Steam  enters  at  the 
steam  inlet  and  passes  through  the  governor  valve  to  the  left 
of  the  first  set  of  blades.  Passing  through  the  alternate  fixed 
and  moving  blades,  it  leaves  through  the  exhaust  outlet. 

Full  peripheral  admission  is  used.  The  increase  in  passage 
area  is  accomplished  by  increasing  the  length  of  blades  from 
stage  to  stage.  The  first-stage  blades  are  very  short,  usually  less 
than  an  inch  in  length.  After  a  large  number  of  stages,  the  blades 
are  of  considerable  length.  To  avoid  further  increase  in  blade 
length,  the  diameter  of  the  drum  is  increased,  and  the  first  blades 
on  the  larger  diameter  are  made  smaller  so  as  to  give  the  proper 
passage  area.  Progressing  to  the  right  the  length  of  blades  again 
increases,  and  again  the  diameter  is  increased.  In  the  last  stage 
the  blades  are  quite  long.  In  the  larger  machines,  the  final  blades 
may  be  as  much  as  a  foot  long. 

Where  the  drum  size  is  increased,  there  is  an  area  exposed  to 
steam  pressure  from  the  left.  This  pressure  causes  an  end-thrust 
on  the  rotor  to  the  right.  To  balance  this  end-thrust  dummy 
or  balance  pistons  are  placed  on  the  left  end  of  the  drum,  each 
one  being  made  of  such  a  diameter  that  the  steam  pressure  on  its 
right  side  produces  the  proper  force  to  the  left.  Initial  steam 
pressure  acts  on  the  right  of  the  smallest  piston  PI.  An  equaliz- 
ing passage,  EI,  leads  from  the  left  of  the  second  stage  on  the 
drum  to  the  right  side  of  the  middle  balance  piston  P2,  so  that  the 
same  steam  pressure  exists  at  6  as  at  c.  In  like  manner,  the  pres- 
sure at  e  is  the  same  as  at  d,  on  the  right  side  of  the  largest  balance 
piston  PS.  A  third  equalizing  pipe,  E$,  connects  the  exhaust  space 
with  the  left  side  of  the  piston  P3.  Of  course  there  is  some  leak- 
age of  steam  by  the  balance  pistons,  but  this  is  minimized  by 
cutting  annular  grooves  in  the  pistons  and  having  rings  on  the 
casing  extend  into  these  grooves  to  form  a  labyrinth  packing. 

Since  the  pressure  drops  in  both  the  fixed  and  movable  blades, 
a  leakage  takes  place,  both  between  the  ends  of  the  fixed  blades 
and  the  drum,  and  over  the  ends  of  the  moving  blades.  It  is 
therefore  essential  that  the  radial  clearance  be  made  as  small  as 
is  safe.  Even  with  the  smallest  clearance  possible,  there  is  bound 


186  ENGINES   AND    BOILERS 

to  be  some  leakage.  The  proportion  of  radial  clearance  space  to 
the  area  of  the  steam  passageway  through  the  blades  is  propor- 
tionally greater  with  the  short  blades  than  with  the  longer  ones. 
Hence  more  leakage  occurs  in  the  first  stage  than  in  the  last. 
It  follows  that  there  is  better  economy  in  a  reaction  turbine  in 
the  low-pressure  stages  than  in  the  high-pressure  stages. 

In  some  makes  of  Parsons  turbines  a  shroud  ring  is  fastened 
over  the  ends  of  the  blades.  In  others,  the  shroud  is  left  off,  but 
the  blades  are  lashed  together  by  means  of  wires  that  pierce  the 
blades.  This  wire  is  comma  shaped  in  cross  section,  and  the  tail 
of  the  comma  is  caulked  down  on  each  side  of  the  blades,  thereby 
keeping  them  in  the  same  relative  position.  These  shroud  rings 
or  lashing  wires  do  not  add  strength  to  resist  the  centrifugal  force, 
but  they  keep  down  the  vibration  or  flutter  of  individual  blades. 
With  long  slender  blades,  the  flutter  might  be  more  than  the  axial 
clearance  and  the  contact  at  high  speed  might  cause  the  blades 
to  be  torn  loose.  With  only  one  blade  loose,  the  whole  system 
of  blading  might  be  almost  instantly  torn  out.  With  this  type 
of  turbine,  it  is  of  the  utmost  importance  that  each  blade  be 
properly  secured  and  adjusted.  With  the  previous  types,  the 
stripping  of  blades  may  be  localized  to  one  stage,  but  in  this, 
the  damage  is  apt  to  be  more  general. 

The  machine  represented  in  Fig.  133  is  equipped  with  an  over- 
load valve.  If  the  load  is  more  than  the  turbine  is  ordinarily 
able  to  carry,  this  valve  is  opened,  allowing  high-pressure  steam 
to  enter  at  the  second  step,  as  shown  by  the  dotted  arrows.  The 
steam  consumption  will  be  greatly  increased,  but  a  much  larger 
load  may  be  carried.  Since  the  governor  still  controls,  the  speed 
may  be  considerably  reduced.  The  overload  valve  is  for  emer- 
gency operation  only,  and  is  not  supposed  to  be  used  often. 

The  two  joints  between  the  shaft  and  the  casing  are  made 
tight  by  a  water  seal.  As  a  turbine  is  ordinarily  run  condensing, 
there  is  a  tendency  for  air  to  leak  in  at  these  joints.  This  leak- 
age of  air  into  the  turbine  is  likely  to  produce  a  greater  loss  of 
efficiency  than  would  a  leakage  of  steam  outward.  The  reason 
for  this  is  that  the  vacuum  in  the  condenser  is  greatly  impaired 
by  the  air  in  the  steam.  If  the  seal  is  kept  full  of  water,  the 
leakage  inward  will  be  of  the  water,  which  will  have  little  or  no 
bad  effect  on  the  economy  of  the  turbine.  In  some  turbines  low- 
pressure  steam  is  used  in  place  of  water  with  the  same  effect. 


STEAM   TURBINES  187 

As  the  weight  of  the  rotor  is  very  great,  the  bearings  B,  B  must 
be  well  constructed  and  kept  cool.  As  in  the  previous  type,  the 
bearings  are  usually  kept  cool  by  means  of  a  circulation  of  oil. 
With  small  clearances,  no  end-play  can  be  allowed.  To  keep 
the  rotor  from  moving  axially,  a  thrust  bearing  is  used.  The  ad- 
justment of  the  thrust  bearing  is  made  by  means  of  the  two 
screws  shown  in  the  figure. 

The  number  of  stages  in  a  reaction  turbine  is  very  great,  which 
necessitates  a  long  rotor.  With  a  long  rotor,  the  changes  in 
length  due  to  temperature  changes  is  considerable.  This  distor- 
tion increases  with  the  amount  of  superheat.  Hence  little  super- 
heat can  be  used  with  some  long  reaction  turbines. 

To  use  more  superheat,  and  also  to  limit  the  length  of  blades 
in  the  later  stages,  the  designers  sometimes  resort  to  a  scheme 
called  compounding,  i.e.,  the  turbine  will  be  cut  into  two  separate 
parts.  The  steam  passes  first  through  the  high-pressure  turbine, 
and  then  is  led  to  the  low-pressure  turbine.  If  the  high-  and 
low-pressure  turbines  are  both  placed  on  the  same  shaft,  the 
machine  is  called  a  tandem-compound  turbine.  Since  they  are 
on  the  same  shaft,  they  must  both  have  the  same  angular  speed. 
Sometimes  better  results  can  be  obtained  by  mounting  each  tur- 
bine on  its  own  shaft  and  running  the  two  at  different  speeds. 
With  the  latter  arrangement,  the  machine  is  called  a  cross-com- 
pound turbine.  In  marine  service,  cross-compound  turbines  are 
sometimes  used,  and  then  each  turbine  is  connected  to  its  own 
propeller  shaft. 

In  order  to  get  rid  of  the  balance  pistons,  Parsons  turbines  are 
sometimes  made  double-flow.  In  the  double-flow  turbine,  the 
steam  enters  at  the  center  of  the  casing  and  half  flows  to  the 
right,  while  the  other  half  flows  to  the  left.  The  two  halves  of 
the  drum  are  exact  duplicates  and  any  end-thrust  on  one  half 
is  balanced  by  the  thrust  on  the  other.  While  this  adds  to  the 
total  number  of  blades  in  the  turbine,  it  does  away  with  the 
dummy  pistons.  Figure  134  shows  the  double-flow  arrangement, 
and  is  self-explanatory.  Quite  often  the  low-pressure  turbine  in 
the  compound  arrangement  is  made  for  double  flow. 

The  governor  used  on  the  Parsons  turbine  made  by  the  West- 
inghouse  Company  is  of  the  blast  type.  As  mentioned  in  §154, 
the  steam  is  admitted  in  blasts  or  puffs.  The  speed  of  the  gov- 
ernor is  much  less  than  that  of  the  shaft,  since  it  is  reduced  by 


188 


ENGINES   AND   BOILERS 


a  worm  gear  from  the  main  shaft.     A  diagram  of  this  type  of 
governor  is  shown  in  Fig.  135.     The  rod  C  is  given  an  up  and 


FIG.  134 

down  motion  from  an  eccentric  on  the  governor  shaft.  The  pivots 
D  and  E  being  fixed,  the  reciprocating  motion  is  communicated 
by  means  of  the  links  and  levers  to  the  small  pilot  valve  A.  The 

recess  on  the  valve  A 
allows  steam  to  enter 
periodically  through 
the  pipe,  to  pass 
through  the  ports,  and 
to  push  up  on  the 
under  side  of  the  pis- 
ton B.  The  fit  around 
the  rod  running  down 
from  B  is  loose,  so 
that  the  pressure  soon 
drops.  The  vertical 

pIG  135  position  of  the  point  F 

is   controlled   by   the 

position  of  the  balls  of  the  governor.  This  in  turn  controls  the 
length  of  time  the  pilot  valve  admits  steam  under  the  piston  B. 


STEAM   TURBINES 


189 


The  piston  B  is  connected  to  the  main  steam  valve  of  the  turbine. 
When  B  is  down,  the  main  steam  valve  is  wide  open.  When  it  is 
up,  the  steam  is  shut  off.  The  length  of  the  time  it  is  up  for  each 
puff  is  seen  to  depend  upon  the  position  of  the  governor  weights. 

161.  The  Westinghouse  Turbine.  —  Aside  from  the  Parsons 
turbine  just  described,  the  Westinghouse  Company  has  made  for 
several  years  a  turbine  which  they  call  the  Westinghouse.  The 
same  type  is  made  under  other  names,  and  has  become  quite 
popular  abroad.  It  consists  of  one  impulse  stage,  such  as  exists 
in  the  Curtis  turbine,  i.e.,  one  pressure  stage,  with  two  velocity 
stages,  and  the  remainder  of  the  turbine  of  the  Parsons  type. 
This  Curtis  stage  may  be  used  with  a  single-flow  Parsons,  or  a 
double-flow  Parsons,  or  with  single-flow  intermediate  Parsons 
stages  combined  with  double-flow  Parsons  in  the  final  stages. 

Figure  136  illustrates  the  Westinghouse  type  in  which  there  is 


FIG.  136 

a  Curtis  stage  combined  with  a  double-flow  Parsons.  The  steam 
enters  the  inlet  at  A  and  passes  through  the  first  stage  exactly 
as  in  the  Curtis  turbine;  it  then  divides,  half  going  to  the  right 
and  half  to  the  left,  through  the  reaction  stages,  and  comes  out 
to  the  exhaust  at  the  two  ends  of  the  casing.  Partial  peripheral 
admission  is  used  in  the  Curtis  stage  and  the  governor  controls 
the  number  of  nozzles  in  use,  as  in  the  Curtis  type. 

There  are  certain  advantages  to  be  gained  by  the  mixed  Westing- 
house  type.     First,  the  length  of  the  rotor  is  shortened,  because 


190  ENGINES   AND   BOILERS 

the  length  taken  up  by  the  Curtis  stage  is  only  a  small  part  of 
that  which  would  be  required  by  the  same  pressure  drop  in  the 
reaction  type.  Second,  the  efficiency  in  the  high-pressure  part 
of  the  turbine  is  increased.  We  have  seen  that  there  is  much 
leakage  in  the  high-pressure  stages  of  the  Parsons  type  because 
the  blade  lengths  are  short  in  these  stages,  and  the  leakage  is 
proportionately  large.  Third,  it  allows  the  governor  to  control 
the  steam  supply  by  shutting  off  nozzles,  which  is  superior  to 
either  the  throttling  or  blast  governing. 

Turbines  are  also  in  use  that  have  one  or  two  Curtis  high-pres- 
sure stages,  and  low-pressure  stages  of  the  Rateau  type.  This 
combination  is  not  common  at  present  in  this  country,  but  it  is 
found  in  some  turbines  in  Europe. 

There  are  many  factors  that  determine  the  choice  of  type  of 
a  turbine.  A  satisfactory  discussion  is  impossible  here.  Given 
the  size,  the  kind  of  service,  and  the  various  operating  conditions, 
the  designer  is  able  to  say  which  type  is  the  best. 

162.  Other  Types.  —  In  addition  to  the  types  heretofore  de- 
scribed, there  are  various  other  small  turbines  in  use.     They  are 
principally  of  the  Pelton  type,  i.e.  they  are  built  on  the  same 
lines  as  a  Pelton  water  wheel.     Buckets  are  cut  in  the  outer 
rim  of  a  rotor  and  the  steam  enters  them  in  a  direction  nearly 
tangential  to  the  rotor.     They  are  usually  built  in  small  sizes 
and  are  used  for  driving  fans,  blowers,  and  pumps.     In  such 
service  simplicity  counts  for  more  than  high  efficiency. 

163.  Low-pressure  Turbines.  —  The  reciprocating  steam  engine 
utilizes   economically  a  greater  amount  of  the  available  energy 
of  the  steam  at  high  pressures  than  at  low  pressures.     That  is, 
the  efficiency  of  a  reciprocating  engine  is  relatively  higher  at  a  prac- 
tical range  of  pressures  above  the  atmosphere  than  below  atmos- 
phere.    This  does  not  mean  that  the  non-condensing  engine  is 
more  efficient  than  the  condensing,  but  that  of  two  engines,  one 
taking  steam  at  a  high  pressure  and  expanding  to  atmospheric 
pressure,  and  the  other  taking  steam  at  atmospheric  pressure 
and  expanding  down  to  that  of  a  good  vacuum,  the  former  will 
be  the  more  efficient.     There  is  about  the  same  amount  of  energy 
available  for  doing  work  in  expanding  from  a  medium  boiler 
pressure  down  to  that  of  the  atmosphere,  as  there  is  in  expand- 
ing from  atmospheric  pressure  to  that  of  a  good  vacuum.     How- 


STEAM   TURBINES  191 

ever,  the  amount  of  expansion  is  much  larger  in  the  second  case.  To 
utilize  this  great  amount  of  expansion  at  low  pressure,  the  cylinder 
would  have  to  be  very  large;  consequently  a  large  cylinder  loss 
would  occur.  With  the  turbine,  the  reverse  is  true,  i.e.  the  turbine 
suffers  relatively  less  loss  at  low  pressures  than  at  high  pressures. 
It  has  been  shown  by  experience  that  the  power  out-put  of  a 
plant  using  non-condensing  engines  can  be  increased  between  80  and 
100  per  cent  by  taking  steam  from  the  engines  and  sending  it 
through  what  are  called  low-pressure  turbines  which  exhaust  into 
a  notably  excellent  vacuum.  This  power  increase  is  obtained 
from  the  steam  without  any  extra  coal  cost  or  any  increase  in 
the  size  of  the  boiler  plant,  through  changing  the  exhaust  pres- 
sure from  that  of  the  atmosphere  to  that  of  the  vacuum  by  the 
introduction  of  the  turbine  and  its  condenser.  It  sometimes 
happens  that  an  existing  plant,  equipped  with  reciprocating  en- 
gines, is  to  be  enlarged,  and  in  some  instances  the  low-pressure 
turbine  has  been  used  to  solve  this  problem.  In  case  the  existing 
engines  are  already  condensing,  it  has  been  found  that  a  net  gain 
in  power  of  25  to  40  per  cent  may  result  by  using  their  exhaust  for 
the  turbines,  on  account  of  the  more  perfect  vacuum  used  with 
turbines  since  they  are  not  subject  to  cylinder  condensation  and 
since  the  air  leakage  is  less.  The  low-pressure  turbine  may  take 
steam  from  engines,  pumps,  air  compressors,  hoists,  etc. 

164.  Mixed-pressure  Turbines.  —  The  mixed-pressure  turbine 
is  much  the  same  as  the  low-pressure  turbine.     It  uses  low- 
pressure  steam,  but  it  may  also  use  some  high-pressure  steam 
at  the  same  time.     The  high-pressure  steam  is  admitted  in  vary- 
ing amounts,  to  make  up  any  deficiency  in  the  supply  of  the  low- 
pressure  steam.     This  may  be  done  by  throttling  the  high-pressure 
steam  down  to  the  low-pressure  before  using  it,  or  the  turbine 
may  be  equipped  with  high-pressure  stages  to  use  the  steam 
without  throttling.     The  latter  is  the  more  efficient  way  of  using 
the  high-pressure  steam. 

165.  Bleeder  Turbines.  —  In  some  plants  equipped  with  tur- 
bines, a  supply  of  low-pressure  steam  is  required  for  heating  or 
for  some  manufacturing  process.     Rather  than  take  high-pressure 
steam  and  throttle  it  down  to  the  required  pressure,  the  low- 
pressure  steam  may  be  drawn  from  the  turbine  at  an  intermediate 
stage.     Then  the  machine  is  said  to  be  a  bleeder  turbine. 


192  ENGINES   AND   BOILERS 

166.  The  Use  of  Superheated  Steam.  —  Turbines  are,  in  gen- 
eral, well  adapted  to  the  use  of  superheated  steam.     Comparative 
tests  show  a  marked  increase  of  efficiency  when  superheated  steam 
is  used.     It  has  been  claimed  that  this  is  due  partly  to  a  lessening 
of  friction  between  the  steam  and  the  parts,  but  it  is  due  mainly 
to  the  increased  available  energy  in  the  steam  which  enters  the 
turbine.     Superheated  steam  also  gives  an  increased  efficiency 
when  used  in  the  reciprocating  engine,  but  its  use  there  is  accom- 
panied by  an  added  difficulty  in  lubrication.     With  the  turbine, 
lubrication  is  no  more  difficult  with  superheated  steam  than  with 
saturated  steam,  since  the  steam-wet  metal  moving  parts  are  not 
in  rubbing    contact  with  other  metal  parts.     Superheating  the 
steam  used  for  turbines  reduces  the  wear  on  the  blading,  com- 
pared with  steam  which  is  saturated  or  wet  at  the  inlet. 

167.  The  Marine  Turbine.  —  The  turbine  has  been  used  in 
marine  service  since  the  early  years  of  its  development.     Many 
factors  make  the  turbine  particularly  well  adapted  to  this  service. 
It  occupies  less  space  than  a  reciprocating  engine,  and  it  is  lighter. 
It  gives  a  uniform  torque  on  the  propeller  shaft  and  does  not 
cause  as  much  vibration  of  the  ship's  hull  as  does  the  recipro- 
cating engine.     On  the  other  hand,  the  turbine  is  a  high-speed 
machine,  while  for  good  efficiency  the  propeller  must  be  run  at 
rather  a  low  speed.     Direct  connection  of  the  turbine  to  the  pro- 
peller shaft  is,  of  course,  the  simplest  arrangement.     When  this 
is  done,  it  is  necessary  to  design  the  turbine  to  run  as  slowly  as 
possible,  and  to  design  the  propeller  to  run  as  fast  as  possible. 
Even  then  a  compromise  often  has  to  be  made:    the  propeller 
has  to  run  too  fast,  and  the  turbine  too  slow,  for  best  efficiency. 
Hence  direct-connected  turbines  are  limited  to  swift  boats.     One 
way  out  of  this  difficulty  is  to  reduce  the  speed  by  means  of  gears. 
While  this  is  sometimes  done,  it  is  not  entirely  satisfactory,  since 
the  gears  are  not  highly  efficient  when  worn. 

A  method  of  drive  that  is  being  used  to  a  considerable  extent 
is  the  electric  drive.  With  this,  turbines  similar  to  those  used 
on  land  are  used  to  drive  electric  generators,  and  the  current 
is  used  by  motors  direct-connected  to  the  propeller  shafts.  This 
allows  both  the  turbine  and  the  propeller  to  run  at  the  proper 
speed.  It  also  permits  great  flexibility  of  arrangement,  and  con- 
venience in  steering  and  in  maneuvering. 


CHAPTER  XI 
GAS  ENGINES 

168.  Introduction.  —  The  small-sized  internal-combustion  en- 
gine is  perhaps  better  known  to  the  average  person  than  any 
other  prime  mover.     During  the  past  twenty  years  it  has  exerted 
a  very  marked  effect  upon  our  manner  of  living.     It  has  made 
possible  the  automobile,  the  motor  truck,  the  gasoline  tractor, 
and  the  airplane.     It  has  replaced  the  more  expensive,  small 
hand  machines  and  horse  machines  on  our  farms.     It  is  indeed 
hard  to  estimate  the  value  to  mankind  of  the  gasoline  engine. 

In  plants  in  which  there  are  combustible  waste  gases,  such  as 
those  from  the  blast  furnaces  and  coke  ovens,  large-size  gas 
engine  units  have  come  into  use,  and  they  are  there  more  eco- 
nomical than  steam  engines.  In  marine  service,  where  space  is 
a  prime  consideration,  as  in  the  submarine,  the  internal-combus- 
tion engine  is  generally  used. 

169.  History.  —  Many  years  ago,  men  of  an  inventive  turn 
of  mind  dreamed  of  gunpowder  engines  or  explosive  engines,  and 
many  patents  were  taken  out  for  these  devices.     Records  show 
that  as  early  as  1680  HUYGHENS  produced  a  working  model  of  a 
gunpowder  engine.     It  was  of  no  practical  importance.     Many 
other  inventors  produced  various  forms  of  engines,  but  not  till 
1860  was  an  internal-combustion  engine  produced  commercially. 
At  that  time  LENOIR  started  building  gas  engines.     In  the  course 
of  a  few  years  four  or  five  hundred  of  these  engines  were  built, 
but  the  engine  was  not  very  efficient  as  it  lacked  compression  for 
the  unignited  gases. 

The  first  really  scientific  work  done  on  the  gas  engine  was  that 
of  the  French  engineer,  BEAU  DE  ROCHAS,  who  laid  down  the  fol- 
lowing four  conditions  as  being  essential  to  high  efficiency. 

(1)  The  largest  cylinder  volume,  with  smallest  exposed  surface, 
i.e.,  the  proper  relation  of  diameter  to  length  of  stroke. 

(2)  The  greatest  possible  rapidity  of  explosion,  i.e.,  the  maxi- 
mum piston  speed. 

(3)  Highest  possible  pressure  at  the  beginning  of  the  expansion. 

(4)  The  greatest  possible  expansion  of  burnt  gases. 

The  same  engineer  proposed  to  obtain  the  above  results  by 
means  of  a  single  cylinder  and  to  operate  his  engine  upon  the 

following  cycle. 

193 


194  ENGINES   AND   BOILERS 

(1)  Draw  in  a  charge  of  mixed  air  and  gas  through  an  entire 
stroke. 

(2)  Compress  this  mixture  during  the  next  stroke. 

(3)  Ignite  the  compressed  combustible  mixture  at  the  beginning 
of  the  third  stroke,  and  expand  the  products  of  combustion  dur- 
ing the  stroke. 

(4)  Discharge  the  burned  gases  on  the  following  stroke. 

The  above  is  known  as  the  four-stroke  cycle,  or  the  Otto  cycle, 
and  is  the  one  most  commonly  used.  It  will  be  considered  more 
in  detail  later. 

A  few  years  after  Beau  de  Rochas  secured  his  patent,  two  in- 
ventors, OTTO  and  LANGEN,  produced  an  engine  that  had  a  ver- 
tical cylinder  with  a  free  piston.  The  explosion  of  uncompressed 
gases  drove  this  piston  upward.  On  its  downward  stroke,  a  rack 
attached  to  the  piston  engaged,  through  a  clutch,  with  a  spur  gear 
that  drove  the  machinery.  While  this  engine  was  more  efficient 
than  any  produced  before,  it  was  noisy.  Although  several  thou- 
sand were  produced,  the  design  was  abandoned  after  a  new  design 
using  compression  of  the  admitted  gases  was  produced  by  Otto. 

The  new  Otto  engine  was  shown  at  the  Paris  Exhibition  in 
1878.  It  operated  on  the  cycle  formulated  by  Beau  de  Rochas, 
and  may  be  considered  as  the  first  modern  gas  engine. 

A  few  years  later,  DUGALD  CLERK  brought  out  an  engine  with  a 
two-stroke  cycle.  Many  machines  of  this  type  are  in  use,  es- 
pecially in  motor  boats  and  the  like. 

Mention  should  be  made  of  the  Brayton  engine,  which  was 
produced  at  about  the  same  time  as  the  Otto-Langen  engine. 
GEORGE  B.  BRAYTON  was  an  American,  and  many  of  his  engines 
were  used  in  this  country.  The  principle  of  its  operation  is  quite 
different  from  that  of  the  others  mentioned,  but  it  need  not  be 
explained  here. 

The  Diesel  engine  dates  back  to  1892,  but  it  was  not  perfected 
until  some  time  later.  Since  then,  the  semi-Diesel  and  other  oil- 
burning  engines  have  come  into  use. 

These  historical  notes  are  given,  not  to  explain  the  principles 
of  operation  of  the  early  engines,  but  to  indicate  the  length  of 
the  period  in  which  the  internal-combustion  engine  was  evolved. 

170.  Cycles  of  Operation.  —  Modern  internal-combustion  en- 
gines operate  upon  either  one  of  two  cycles:  the  four-stroke 


GAS   ENGINES 


195 


cycle  or  the  two-stroke  cycle.  These  terms  are  usually  abbre- 
viated to  four-cycle  and  two-cycle.  The  four-stroke  cycle  is  some- 
times called  the  Otto  cycle,  and  the  two-stroke  cycle  is  occa- 
sionally known  as  the  Clerk  cycle.  Each  of  these  cycles  is  used 
with  gas,  gasoline,  Diesel,  semi-Diesel,  and  other  oil  engines. 

171.  The  Four-stroke  Cycle.  —  In  the  four-stroke  cycle  there 
are  four  strokes  of  the  piston,  two  forward,  and  two  backward. 
These  occur  as  the  shaft  makes  two  complete  revolutions.  Figure 


Met 


fxhoust 


FIG.  137 


137  represents  an  engine  cylinder  with  its  piston  P.  The  inlet 
valve  is  at  A,  and  the  exhaust  valve  at  B.  When  the  piston  is 
at  its  crank-end  dead-center  position,  the  volume  to  the  left  of 
it  is  the  piston  displacement  plus  the  clearance.  Both  valves,  A 
and  B,  are  closed,  and  the  space  to  the  left  of  the  piston  is  filled 
with  a  mixture  of  air  and  fuel. 

During  the  first  stroke,  the  piston  moves  from  its  crank-end  to 
its  head-end  dead  center,  compressing  the  mixture  into  the  clear- 
ance space.  Near  the  head-end  dead  center,  the  compressed  mix- 
ture is  ignited,  whereupon  it  burns,  and  its  pressure  suddenly 
increases. 

On  the  second  stroke  the  piston  moves  from  the  head-end  dead 
center  to  the  crank-end  dead  center,  while  the  burnt  gases  expand 
and  do  work  upon  the  piston.  Near  the  end  of  the  second  stroke, 
the  exhaust  valve  B  opens. 

As  the  piston  moves  to  the  left  on  the  third  stroke,  the  burnt 
gas  is  forced  out  to  the  exhaust.  The  burnt  gas  that  is  left  in 
the  clearance  space  is  not  expelled.  At  the  end  of  the  third 
stroke  the  exhaust  valve  closes. 

At  the  beginning  of  the  fourth  stroke,  the  inlet  valve  A  opens. 
As  the  piston  moves  to  the  right,  a  fresh  charge  of  air  and  fuel 


196 


ENGINES   AND    BOILERS 


is  sucked  into  the  cylinder.     At  the  end  of  the  fourth  stroke  the 
inlet  valve  closes.     This  completes  the  cycle  of  operation. 

Figure  138  shows  the  indicator  card  of  the  four-stroke  cycle 
engine.     Starting  from  E,  at  a  little  below  the  atmospheric  pres- 


Atmospheric  Line 

FIG.  138 

sure,  the  charge  is  compressed  during  the  first  stroke  to  the 
pressure  at  the  point  A.  From  A  to  B,  the  charge  is  burned 
and  the  pressure  rises.  From  B  to  C  expansion  takes  place.  From 
C  to  D  the  burnt  gases  are  expelled.  The  suction  stroke  is  rep- 
resented by  DE.  The  four  strokes  then  represent  compression, 
burning  and  expansion,  scavenging,  and  suction.  On  the  card,  L 
represents  the  length  of  stroke,  and  F  the  clearance. 

172.  The  Two-stroke  Cycle.  —  This  cycle  is  completed  in  one 
revolution.     Figures  139  and  140  show  a  two-stroke  cycle  engine 


FIG.  139 

such  as  is  often  used  on  motor  boats.  Due  to  the  fact  that  the 
piston  covers  and  uncovers  the  admission  and  the  exhaust  ports, 
it  is  often  called  a  valueless  engine. 


GAS   ENGINES 


197 


In  Fig.  139,  the  piston  is  shown  at  the  crank-end  dead  center. 
Both  the  inlet  and  exhaust  ports  are  open.  There  is  a  small 
compression  in  the  crank  case  and  the  combustible  mixture  of 
air  and  fuel  is  forced  up  and  into  the  cylinder  through  the  port  A. 
There  is  a  baffle  on  the  top  of  the  piston  which  deflects  the  in- 
coming mixture  to  the  top  of  the  cylinder.  At  the  same  time  the 
burnt  gases  from  the  previous  stroke  are  escaping  through  the 


FIG.  140 

exhaust  port  B.  Naturally  a  little  of  the  unburned  gas  will  es- 
cape before  the  exhaust  port  is  closed. 

As  the  piston  moves  upward  on  its  first  stroke,  it  covers  the 
ports  A  and  B,  and  then  compresses  the  mixture  into  the  clear- 
ance space.  At  the  head-end  dead  center,  Fig.  140,  ignition  takes 
place,  and  the  piston  is  forced  downward  on  the  second  stroke 
by  the  pressure  produced.  The  burnt  gases  expand  until  the 
exhaust  port  B  is  uncovered.  They  then  escape  to  the  exhaust. 
As  the  piston  moves  on  down,  the  inlet  port  A  is  uncovered,  and 
the  fresh  gas  coming  in  at  A  sweeps  out  more  of  the  burnt  gas. 
As  the  piston  moves  upward  a  slight  vacuum  is  formed  in  the 
crank  case.  When  the  piston  gets  nearly  to  the  top  of  its  travel, 
a  port  communicating  to  the  carburetor  or  fuel  and  air  supply 
is  uncovered,  and  the  combustible  mixture  is  sucked  into  the 
crank  case.  Upon  the  piston's  downward  stroke  this  mixture  is 
compressed  enough  to  force  it  into  the  cylinder  when  the  port  A 
is  uncovered. 

In  Fig.  141  another  two-stroke  cycle  engine  is  shown.  In  the 
large  size  of  these  engines,  separate  pumps  for  air  and  gas  are 


198 


ENGINES   AND   BOILERS 


used  instead  of  the  compression  in  the  crank  case.  The  engine 
is  double-acting,  and  the  exhaust  port  is  placed  around  the  cylinder 
midway  between  the  two  ends.  The  piston  P  uncovers  this  ex- 
haust port  near  the  end  of  each  stroke.  Gas  and  air  is  com- 
pressed in  the  pumps  shown.  The  piston  valves  of  the  pumps 
deliver  the  gas  and  air  alternately  to  the  two  ends  of  the  cylinder. 


FIG.  141 

The  air  and  gas  are  mixed  just  before  they  enter  the  admission 
valve  I.  In  Fig.  141,  gas  and  air  are  compressed  in  the  left  ends 
of  the  pumps,  and  are  forced  into  the  left  end  of  the  engine  cylinder. 
As  the  piston  starts  to  the  left,  the  left  admission  valve  closes, 
and  the  mixture  is  compressed  into  the  clearance  space.  At  dead 
center  the  charge  is  fired,  and  expansion  takes  place  as  the  piston 
moves  to  the  right.  When  the  piston  uncovers  the  exhaust  port, 
the  burnt  gases  escape.  In  Fig.  141,  the  amount  of  gas  forced 
into  the  cylinder  is  controlled  by  butterfly  valves  whose  position 
is  controlled  by  the  governor. 

The  indicator  card  of  the  two-stroke  cycle  engine  of  Figs.  139 
and  140  is  shown  in  Fig.  142.  Compression  takes  place  from 
E  to  A,  burning  from  A  to  B,  and  expansion  from  B  to  C.  From 
C  to  D,  and  from  D  to  E,  the  'exhaust  of  the  burnt  gases  takes 
place.  The  admission  of  the  charge  occurs  from  H  to  D  and 


GAS   ENGINES 


199 


from  D  to  G.     It  is  seen  that  the  exhaust  port  opens  a  little 
sooner  than  the  inlet  port. 

173.  Classification  from  Fuel  Used.  —  Internal-combustion  en- 
gines are  called  by  different  names,  according  to  the  fuel  used. 
We  have  the  gas,  gasoline,  oil,  Diesel  and  semi-Diesel  engines. 
The  fundamental  principle  is  much  the  same  in  all  of  them,  dif- 
ference being  largely  a  matter  of  detail  in  design  and  in  the  feed 
of  the  fuels. 

In  the  earlier  gas  engines,  city  or  coal  gas  was  often  used. 
Then,  in  the  days  of  the  natural  gas  booms  in  this  country,  gas 
engines  became  quite  common.  With  these  kinds  of  gas,  it  was 


-&         Atmospheric  £/ne 

FIG.  142 

not  possible  to  give  the  mixture  of  fuel  and  air  a  very  high  com- 
pression, because  the  temperature  during  compression  might 
be  so  high  that  ignition  would  occur  before  the  end  of  the  com- 
pression stroke.  With  the  use  of  producer  gas  or  blast-furnace 
gas  the  compression  can  safely  be  carried  higher;  hence  we  find 
it  common  to  use  a  less  amount  of  clearance  with  engines  designed 
to  use  a  lean  gas  for  fuel.  With  a  small  clearance  the  compres- 
sion is  higher. 

The  internal  combustion  engine  with  which  we  are  most  familiar 
is  the  gasoline  engine.  Gasoline  will  vaporize  partially  at  ordi- 
nary temperatures;  hence  a  spray  of  the  liquid  fuel  is  mixed 
with  the  air  as  it  goes  to  the  cylinder.  While  this  spray  is  com- 
monly not  all  vaporized  before  reaching  the  cylinder,  it  is  sufficiently 
vaporized  so  that  an  explosive  mixture  results,  and  the  charge 
is  fired.  If  the  spray  is  fine  enough  and  if  sufficient  time  is  given 
during  the  stroke,  the  gasoline  will  be  very  nearly  all  buined. 
Of  course  there  are  differences  in  gasolines,  some  kinds  being  more 
volatile  than  others.  The  device  that  introduces  the  spray  into  the 


200 


ENGINES   AND    BOILERS 


air  intake  is  called  a  mixing  valve,  or  a  carburetor.  With  the  less 
volatile  liquid  fuels,  such  as  kerosene,  heat  is  sometimes  applied 
to  help  in  vaporizing  it  before  it  is  introduced  into  the  cylinder. 

When  liquid  fuels  heavier  than  gasoline  are  used,  it  is  common 
to  spray  them  into  the  cylinder  during  the  compression  stroke. 
Often  the  spray  strikes  a  hot  plate  in  the  cylinder  and  the  fuel  is 
sufficiently  vaporized  so  that  ignition  can  take  place  at  the  end  of 
the  stroke.  The  compression  in  low-pressure  oil  engines  is  no 
greater  than  in  some  gas  engines,  usually  about  60  pounds  per 
square  inch. 

In  the  Diesel  engine  the  clearance  is  much  less  and  a  com- 
pression of  500  to  600  pounds  per  square  inch  is  attained.  Prac- 


df/nospheric  Line 
FIG.  143 

tically  all  liquid  fuels  will  partially  burn  at  a  temperature  at- 
tained by  the  compression  of  air  to  200  to  300  pounds  per  square 
inch.  To  prevent  premature  ignition,  the  liquid  fuel  is  sprayed 
into  the  cylinder  at  the  beginning  of  the  forward  or  working 
stroke  in  Diesel  engines. 

Air  alone  is  compressed  during  the  compression  stroke.  At  a 
pressure  of  500  to  600  pounds  its  temperature  will  be  in  the 
neighborhood  of  1000°  F.  This  is  sufficient  to  ignite  and  com- 
pletely burn  the  oil  that  is  sprayed  in.  The  length  of  time  the 
oil  is  injected  into  the  cylinder  usually  is  regulated  by  the  gov- 
ernor. The  Diesel  engine  may  operate  on  either  the  four-stroke 
or  on  the  two-stroke  cycle. 

The  indicator  diagram  for  the  four-cycle  Diesel  engine  is  shown 
in  Fig.  143.  On  the  first  stroke,  air  is  compressed  from  a  pressure 
a  little  below  that  of  the  atmosphere,  shown  at  E,  to  500  or  600 


GAS   ENGINES 


201 


pounds  per  square  inch,  as  shown  at  A.  From  A  to  B,  the  fuel 
is  injected  and  burned.  Expansion  of  the  burnt  gas  takes  place 
from  B  to  C.  The  cylinder  is  scavenged  from  C  to  D,  and  a 
fresh  charge  of  air  is  drawn  in  from  D  to  E. 

The  high  pressures  necessary  in  the  Diesel  engine  have  been 
found  troublesome  from  a  mechanical  standpoint,  and  there  has 
been  a  tendency  to  reduce  the  high  compression.  With  a  com- 
pression around  200  to  300  pounds  the  term  semi-Diesel  is  used. 
There  is  but  little  difference  in  the  theory  of  operation  between 
the  Diesel  and  semi-Diesel.  Oil  is  injected  at  the  opening  of 
the  expansion  stroke  as  before.  However,  on  account  of  the 


FIG.  144 

lower  temperature  of  compression,  aid  to  the  vaporization  of  the 
oil  is  given  by  the  addition  of  a  hot  bulb  located  in  the  head 
of  the  cylinder. 

Figure  144  shows  this  type  of  engine.  Before  starting,  the 
bulb  is  heated  by  means  of  the  burner  B.  After  the  engine  has 
started  the  bulb  will  be  hot  enough  without  the  aid  of  the  burner. 
In  Fig.  144,  air  is  drawn  into  the  crank  case  and  compressed  as 
the  piston  moves  forward.  With  the  piston  near  the  crank-end 
dead  center,  the  exhaust  port  is  uncovered  and  the  air  is  blown 
in  through  the  ports  at  the  top  of  the  cylinder.  As  the  piston 
returns  to  the  left,  the  charge  of  fresh  air  is  compressed.  With 


202  ENGINES   AND    BOILERS 

the  piston  on  head-end  dead  center,  the  oil  pump  injects  the  fuel 
into  the  cylinder,  and  the  spray  strikes  the  lip  of  the  hot  bulb. 
The  temperature  of  the  bulb  is  high  enough  so  that  the  fuel  is 
ignited  and  practically  all  burned. 

This  engine  operates  on  the  two-stroke  cycle.  On  the  larger  en- 
gines the  four-stroke  cycle  is  more  common.  On  heavy  loads, 
there  is  often  a  tendency  to  knock  in  the  semi-Diesel  engine.  This 
is  relieved  by  the  injection  of  a  small  amount  of  water  with  the 
fuel,  which  does  not  seem  to  lessen  the  efficiency  of  the  engine. 

174.  Efficiency.  —  The  efficiency  of  an  engine  operating  on  the 
Carnot  cycle  is  (T\—Tz)/Ti,  where  T\  is  the  absolute  tempera- 
ture during  combustion,  and  T^  the  absolute  temperature  of  the 
gases  in  the  exhaust.     Of  course  none  of  our  engines  operate  on 
the  Carnot  cycle,  but  their  efficiency  does  depend  upon  the  range 
of  temperature  in  the  cylinder.     Other  factors  being  the  same,  it  is 
evident  that  the  highest  thermal  efficiency  will  occur  in  that  engine 
which  has  the  largest  range  of  temperatures  during  the  working 
stroke.     From  this  it  is  seen  that  those  engines  with  the  highest 
compressions,  and  therefore  the  highest  combustion  temperatures, 
will   have   the   highest   efficiency.     This  also   explains  the  fact 
that  gas  engines  using  a  lean  gas,  such  as  blast-furnace  gas,  may 
have  a  higher  efficiency  than  those  that  operate  on  natural  gas 
or  gasoline  vapor.     Of  all  the  internal-combustion  engines,  the 
Diesel  has  the  highest  thermal  efficiency.     The  efficiency  based 
on  the  brake  horsepower  ranges  from  30  to  35  per  cent.     Other 
gas  engines  give  somewhat  lower  efficiency. 

175.  Fuels.  —  Gas  engines  may  operate  on  almost  any  com- 
bustible gas.     Naturally  the  more  expensive  gases  are  but  little 
used.     Ordinary  city  gas  is  commonly  a  mixture  of  coal  gas  and 
water  gas.     As  ordinarily  produced  it  is  too  expensive  for  exten- 
sive use  in  gas  engines.     Where  natural  gas  is  plentiful  and  cheap, 
it  is  commonly  used  for  gas  engines. 

At  some  blast-furnace  plants,  the  gas  from  the  furnace  is  used 
for  fuel  in  the  gas  engines  that  drive  the  blowers.  This  gas  is 
not  of  what  we  commonly  call  high  quality,  that  is  its  heat  con- 
tent per  cubic  foot  is  much  lower  than  that  of  natural  gas  or 
coal  gas,  but  it  gives  excellent  results  when  used  in  the  engines. 

The  by-product  gas  from  coke  ovens  is  being  used  more  and 
more  as  the  efficiency  of  these  plants  is  looked  after.  There  are 


GAS   ENGINES  203 

also  quite  a  number  of  plants,  especially  in  the  east,  where  pro- 
ducer gas  is  used  for  gas  engines. 

Natural  gas,  while  it  varies  in  composition,  may  be  said  to  be 
composed  mainly  of  marsh  gas,  CH4.  In  some  natural  gases, 
there  is  a  considerable  free  hydrogen  and  also  an  appreciable 
amount  of  olefiant  gas,  C2H4.  The  heat  value  of  natural  gas  usu- 
ally is  between  900  and  1000  B.t.u.  per  cubic  foot. 

The  illuminating  gas  used  in  cities  varies  widely  in  its  com- 
position, depending  upon  how  it  is  made.  It  usually  contains 
about  40  per  cent  H2,  30  per  cent  CH4,  and  varying  amounts  of 
CO  and  C2H4  besides  CO2  and  N2.  The  heating  value  aver- 
ages from  500  to  600  B.t.u.  per  cubic  foot. 

Coke-oven  gas  contains  about  50  per  cent  H2  and  35  per  cent 
CH4.  Its  heating  value  is  about  the  same  as  that  of  illuminating 
gas. 

The  principal  combustible  substance  in  blast  furnace  gas  is  CO, 
which  ordinarily  runs  about  25  per  cent.  The  heating  value  of 
blast-furnace  gas  is  but  little  over  100  B.t.u.  per  cubic  foot. 

Producer  gas,  while  it  is  variable,  contains  about  15  per  cent 
H2,  and  25  per  cent  CO.  It  has  a  heat  value  of  about  145  B.t.u. 
per  cubic  foot. 

Practically  all  these  gases  contain  a  little  02  and  varying  amounts 
of  C02  and  N2.  These  substances  in  the  gas  add  no  heating  value 
to  it. 

The  liquid  fuels  used  in  internal-combustion  engines  vary  from 
crude  oil  to  more  refined  products,  such  as  gasoline.  Many 
Diesel  engines  seem  able  to  burn  crude  oil  very  well,  but  some 
of  the  semi-Diesel  and  oil  engines  do  better  on  the  more  volatile 
product,  such  as  kerosene.  The  great  demand  for  gasoline  has 
led  to  a  gradual  lowering  of  the  flash-point  of  this  product.  Car- 
buretors that  used  to  give  good  results  with  the  gasoline  sold  a 
few  years  ago,  now  have  trouble  in  using  the  commercial  grades. 
The  development  of  carburetors  has  had  to  keep  step  with  the 
change  in  the  quality  of  the  product  they  have  to  handle. 

176.  The  Gasoline  Carburetor.  —  There  are  a  great  variety  of 
mixing  valves  or  carburetors  on  the  market.  We  cannot  hope 
to  describe  all  of  the  various  types  in  this  course.  Only  one  simple 
form  will  be  described,  but  the  fundamental  principle  is  the  same 
in  all.  This  principle  is  to  divide  the  liquid  fuel  into  as  fine  a 


204 


ENGINES   AND   BOILERS 


spray  as  possible  and  to  mix  it  thoroughly  with  air.  Some  of 
the  liquid  is  evaporated,  but  it  is  doubtful  whether  it  is  ever  all 
vaporized. 

Evaporation  is  not  necessary  if  the  liquid  particles  are  finely 
enough  divided  and  are  thoroughly  mixed  with  the  air.  Even 
solid  combustible  matter  is  explosive  when  mixed  with  air,  as 
is  shown  by  flour-mill  explosions  and  the  explosions  in  mines  due 
to  dust  of  inflammable  materials. 

Figure  145  shows  a  gasoline  carburetor.  The  main  body  of 
the  carburetor  B  is  partly  filled  with  gasoline.  The  height  of 


FIG.  145 

the  liquid  is  kept  very  near  to  a  constant  level  by  means  of  a 
float  F,  which  is  attached  by  means  of  a  lever  to  the  valve  H 
which  leads  from  the  supply.  Air  enters  through  the  opening 
at  the  top  and  passes  down  through  the  passage  C.  Turning  to 
the  left,  it  passes  out  to  the  pipe  leading  to  the  engine.  A  spray 
of  liquid  is  injected  into  the  air  through  the  needle  valve  D. 
The  opening  in  the  needle  valve  D  is  adjusted  by  the  handle  E. 
The  opening  of  the  needle  valve  into  the  air  passage  is  a  little 
above  the  liquid  level  in  B,  so  that  no  gasoline  runs  out  unless 
there  is  a  current  of  air  to  suck  up  the  liquid.  There  is  at  all 
times  an  opening  for  the  air  to  enter  the  carburetor  at  A,  but 
this  opening  may  be  increased  when  a  large  supply  is  needed  by 
the  suction  pulling  back  the  valve  A.  The  valve  A  is  held  on 
its  seat  by  a  light  spring  0.  The  tension  in  the  spring  0  may  be 


GAS   ENGINES  205 

adjusted  by  turning  the  screw  M.  By  the  proper  adjustment 
of  the  air  valve  A  and  the  needle  valve  Z),  any  richness  of  mix- 
ture may  be  secured.  The  amount  of  mixture  of  air  and  gaso- 
line leaving  the  carburetor  to  enter  the  engine  is  regulated  by  the 
throttle  K.  The  bowl  of  the  carburetor  may  be  drained  by  open- 
ing the  cock  T. 

177.  The  Gas  Producer.  —  A  large  amount  of  publicity  has 
been  given  to  the  subject  of  gas  producers  for  the  gas  engine, 
and  much  research  work  has  been  done  on  the  subject.  It  is  only 
just,  however,  to  say  that  the  producer  plant  has  been  somewhat 
of  a  disappointment  in  America.  While  trial  tests  show  up  re- 
markably well,  and  the  claims  of  makers  are  unusually  good, 
actual  experience  has  shown  that  the  time  has  not  yet  come  when 
they  can  replace  the  steam  plant.  It  is  not  safe  to  predict  as 
to  the  future,  and  every  power  plant  engineer  should  be  some- 
what acquainted  with  the  subject. 

When  air  is  passed  through  a  bed  of  hot  carbon,  combustion 
takes  place.  If  there  is  sufficient  air,  the  combustion  is  complete 
and  CC>2  is  formed.  If  not  enough  air  is  supplied,  the  burning 
is  only  partial  and  CO  is  formed.  This  CO  may  later  be  burned 
to  CO2  by  the  addition  of  more  air,  which  is  done  in  the  engine 
cylinder  in  the  case  of  the  producer  and  engine  plant. 

If  steam  is  passed  through  a  hot  carbon  bed,  a  decomposition 
of  the  steam  takes  place.  The  hydrogen  is  liberated  as  H2  and  the 
oxygen  combines  with  the  carbon  to  form  CO.  Both  of  these 
gases  are  valuable  as  fuel,  and  the  mixture  is  often  called  water 
gas. 

When  air  is  passed  through  the  hot  carbon  bed,  and  CO  is 
formed,  heat  is  generated,  so  that  the  bed  gets  hotter  and  hotter. 
On  the  other  hand,  when  steam  is  passed  through,  heat  is  ab- 
sorbed and  the  bed  gets  cooler  and  cooler.  By  the  proper  pro- 
portioning of  air  and  steam  passed  through,  it  is  possible  to  keep 
the  fuel  bed  at  the  proper  temperature.  This  is  what  is  done 
in  the  gas  producer.  It  is  evident  that  most  of  the  gases  in 
producer  gas  will  be  CO,  H  and  N,  the  nitrogen  coming  from  the 
air  and  being  inert. 

The  fuel  used  in  the  producer  may  be  coke  or  coal.  Better 
results  are  obtained  by  using  anthracite  coal  than  by  using  bitu- 
minous coal.  This  is  partly  due  to  the  fact  that  bituminous  coal 


206 


ENGINES   AND   BOILERS 


tends  to  cake  and  needs  constant  working  or  poking  to  keep  holes 
from  burning  through  the  cake,  thereby  letting  excess  air  get 
through.  Moreover,  bituminous  coal  gives  off  various  tars  when 
it  is  heated.  If  these  are  not  removed,  they  clog  up  the  pipes 
and  the  engine.  The  removal  of  the  tar  is  not  easy.  It  is  some- 
times done  by  throwing  the  tarry  material  out  of  the  gas  by 
centrifugal  force  by  means  of  a  kind  of  fan  arrangement,  or  by 
passing  the  gas  through  scrubbers.  Devices  have  been  tried 


FIG.  146 

whereby  the  distilled  products  are  passed  through  the  hot  fuel 
bed  and  their  composition  thereby  changed.  This  last  method 
resembles  the  underfeed  furnace  used  in  steam  plants. 

Figure  146  shows  a  form  of  gas  producer.  Coal  is  fed  into  the 
hopper  A.  From  this,  it  is  dropped  into  the  chamber  B.  Pass- 
ing out  of  the  bottom  of  B,  it  is  scattered  in  a  uniform  layer  over 
the  fuel  bed  by  means  of  the  spiral  spreader  C,  which  is  rotated 
by  the  bevel  gear  Q. 

The  fuel  bed  may  be  divided  roughly  into  three  zones.  The 
top  zone  is  the  green-coal  zone,  where  distillation  takes  place. 
The  volatile  products  pass  on  out  with  the  other  gas. 

As  the  volatile  products  are  driven  off,  and  the  bed  settles,  the 
fuel  reaches  the  coke  zone.  It  is  here  that  the  burning  and  de- 
composition mentioned  above  take  place.  As  the  carbon  is  burnt 


GAS   ENGINES  207 

out,  the  ash  settles  to  the  bottom  of  the  producer,  where  it  is 
raked  out  through  the  water  seal  R. 

Air  from  the  duct  E  is  led  to  the  bottom  of  the  fuel  bed  through 
D,  and  passes  up  through  the  coke  zone  and  the  green-coal  zone. 
Steam  is  admitted  to  the  air  supply  through  the  pipe  G.  Holes 
are  provided  in  the  furnace  walls  at  F  for  the  working  of  the  fuel 
bed. 

The  gas  leaves  the  producer  through  the  opening  at  7,  and 
passes  down  the  pipe  shown  to  K.  From  K  it  passes  through 
the  pipe  L  to  the  wet  scrubber.  The  wet  scrubber  is  filled  with 
coke,  Pj  which  is  continuously  sprinkled  with  water  from  the 
nozzle  N.  The  gas  passing  up  through  the  wet  coke  is  cooled 
and  deposits  dust,  tar  and  other  impurities.  The  gas  leaves  the 
wet  scrubber  at  0,  and  may  go  either  directly  to  the  engine  or 
else  to  a  dry  scrubber.  The  dry  scrubber  is  filled  with  wood 
shavings  or  excelsior,  which  takes  out  the  remaining  tar. 

Upon  starting  the  producer,  the  gas  is  vented  to  the  roof  through 
the  valve  J.  As  soon  as  the  quality  of  the  gas  becomes  good 
enough,  the  valve  J  is  closed  and  the  engine  is  started.  The  coal 
in  B  is  kept  cool  enough  to  prevent  it  from  burning,  by  a  water- 
jacket  S.  The  producer  is  lined  with  firebrick. 

Air  either  may  be  blown  into  the  producer,  or  it  may  be  drawn 
through  by  the  suction  of  the  engine.  In  the  former  case  a 
storage  tank  for  the  gas  is  necessary.  With  the  suction  type  the 
storage  is  unnecessary,  as  the  engine  draws  through  only  what  it 
needs.  When  a  water  seal  is  used  at  the  bottom,  the  producer 
is  called  a  wet  bottom  producer.  The  poking  of  the  fuel  bed 
may  be  done  by  hand  or  mechanically. 

178.  Cooling  of  Cylinders.  —  The  cylinder  walls  of  the  internal- 
combustion  engine  must  be  kept  cool  enough  to  insure  proper 
lubrication.  This  cooling  is  commonly  done  by  circulating  water 
through  a  jacket  around  the  cylinder,  as  shown  at  W,  in  Figs. 
137,  139,  140,  141  and  144.  As  far  as  the  efficiency  of  the  engine 
is  concerned,  the  hotter  the  cylinder  walls  the  better,  so  that  it 
is  evident  that  they  should  not  be  cooled  any  more  than  is  neces- 
sary to  insure  lubrication.  In  small  engines,  the  cylinder  is 
sometimes  placed  in  the  lower  part  of  a  hopper  filled  with  water. 
Fresh  water  is  added  as  it  boils  away  in  the  hopper. 

Another  method  that  is  used  occasionally  to  cool  the  cylinder 


208 


ENGINES   AND   BOILERS 


Met 


FIG.  147 


is  by  having  the  outer  walls  of  the  cylinder  and  head  covered  with 
fins.  These  present  a  large  surface  to  the  air  and  the  heat  is 
radiated  from  them.  To  increase  this  transfer  of  heat,  a  cur- 
rent of  air  is  kept  moving  over  the 
surface  of  the  fins  by  means  of  a 
fan.  This  latter  method  is  called 
air-cooling.  Figure  147  shows  an 
air-cooled  cylinder. 

179.  Ignition.  —  The  charge  of 
combustible  mixture  in  an  internal- 
combustion  engine  is  fired  in  vari- 
ous ways.  A  method  formerly  used 
quite  extensively  but  now  not  very 
common  is  the  hot-tube  method. 
This  method  is  illustrated  by  the 
sketch  in  Fig.  148.  The  tube  which 
connects  with  the  cylinder  is  heated 
by  a  gas  jet.  Arrangement  is  made 

so  that  the  flame  can  be  shifted  along  the  tube,  heating  it  closer 

or  farther  away  from  the  cylinder. 

To  explain  how  the  scheme  works,  suppose  the  charge  has  been 

fired.     The  tube  will  then  be  filled  with  burnt  gas.      After  the 

fresh  charge  has  been  drawn  in 

and  compression  started,  the  fresh 

gas  is  forced  into  the  tube.    When 

it  is  compressed  into  the  tube  far 

enough  to  strike  the  heated  part, 

ignition    occurs.     Naturally,   this 

scheme  can  be  used  only  when  the 

load  is  constant. 

As    has    been   explained  previ- 
ously,   ignition   may   be   had    by 

using  a  high  compression,  so  that 

the   temperature    of    compression 

will  be  high  enough  to   fire  the 

charge.      This    method    is    used 

mostly  in  oil  engines  and  is  assisted  in  those  using  the  lower 

pressures  by  a  hot  bulb  or  plate.     The  hot  bulb  acts  somewhat 

in  the  same  manner  as  the  hot  tube  just  mentioned.     With  gas 


FIG.  148 


GAS   ENGINES  209 

or  the  more  volatile  liquid  fuels  this  method  is  not  satisfactory, 
since  early  ignition  is  apt  to  occur. 

The  most  common  method  is  electric  ignition.  There  are  two 
general  types  of  electric  ignitors,  the  jump-spark  and  the  make- 
and-break.  In  the  former,  a  spark  plug  is  used  (Fig.  149),  which 
has  two  fixed  terminals  exposed  to  the  gases  of  the  cylinder. 
At  the  proper  time  for  ignition  a  current  with  voltage  high  enough 
to  jump  the  gap  is  introduced  in  the  circuit.  The 
heat  of  this  spark  ignites  the  charge.  The  details  of 
timing  and  of  producing  the  current  will  not  be  dis- 
cussed here,  except  to  say  that  the  current  may  be 
furnished  either  by  dry-cell  or  storage  batteries,  or  by 
a  magneto. 

In  the  make-and-break  system,  two  electrodes  are 
brought  into  contact  within  the  cylinder  and  are  sepa- 
rated at  the  proper  time  for  ignition.  As  the  circuit  is 
broken  a  spark  is  formed  between  them.  The  details 
of  the  many  schemes  used  will  not  be  discussed  here. 
This  make-and-break  system  does  not  require  as  high 
an  e.m.f.  as  the  jump-spark  system,  but  it  is  limited  to  the  slower 
engine  speeds. 

180.  Valves.  —  The  earlier  types  of  gas  engines  were  equipped 
with  slide-valves.      These  required  lubrication,  which  is  some- 
what difficult  at  high  temperatures.     Except  for  a  few  engines 
that  use  sleeve-valves,  gas  engines  of  today  have  the  so-called 
lifting  or  poppet  valves ,  which  are  commonly  lifted  .by  cams  on 
a  cam  shaft.     In  the  four-stroke  cycle  engines  the  cam  shaft  is 
geared  to  run  at  half  the  speed  of  the  crank  shaft,  so  that  the 
cam  shaft  makes  one  revolution  per  cycle.     The  cams  are  placed 
on  the  cam  shaft  so  that  the  valves  open  and  close  at  the  proper 
time.     Each  valve  is  kept  seated  by  means  of  a  spring  around 
the  valve  stem.     Figure  147  shows  the  cams,  and  how  they  are 
made  to  lift  the  valves.     In  some  of  the  slower  speed  engines, 
the  inlet  valve  is  not  opened  by  means  of  a  cam,  but  by  the 
suction  inside  the  cylinder.     With  •  this  arrangement  a    strong 
spring  cannot  be  used  to  seat  the  valve. 

181.  Governing.  —  As  far  as  the  governor  itself  is  concerned, 
the  gas-engine  governor  does  not  differ  materially  from  the  steam- 
engine  governor.     Both  centrifugal  and  inertia  governors  are  used. 


210  ENGINES   AND   BOILERS 

Depending  upon  how  the  governor  regulates  the  amount  of  work 
done  in  the  engine  cylinder,  we  have  three  general  types  of  gas 
engine  governors:  (1)  hit-and-miss  governors,  (2)  quantity  gov- 
ernors, and  (3)  quality  governors. 

HIT-AND-MISS  GOVERNOR.  With  this  type  of  governor  there 
is  a  working  stroke  for  every  cycle  under  conditions  of  maximum 
load.  At  lighter  loads  the  governor  mechanism  fails  to  admit 
a  charge  occasionally,  giving  what  might  be  termed  a  blank  cycle 
in  which  no  work  is  done  by  the  cylinder.  This  drops  the  speed 
of  the  engine  and  the  governor  acts  so  that  fuel  is  again  taken  in 
as  before.  With  this  scheme  the  engine  either  operates  under 
conditions  of  maximum  efficiency,  or  it  does  not  fire  at  all.  This 
method  of  governing  gives  better  economy  at  light  loads  than  the 
other  methods,  but  it  does  not  give  close  speed  regulation.  When 
the  engine  misses,  the  exhaust  valve  is  commonly  held  open  so 
that  there  is  no  work  done  in  useless  compression. 

QUANTITY  GOVERNOR.  For  every  engine  there  is  a  ratio  of  gas 
to  air,  which  is  nearly  constant,  with  which  the  engine  gives  the 
best  efficiency.  With  the  quantity  governor,  this  ratio  is  kept 
(theoretically)  constant.  Regulation  is  accomplished  in  two  ways : 
(1)  by  the  cut-off  governor,  and  (2)  by  the  throttling  governor. 
With  the  cut-off  or  throttling  governor,  the  normal  mixture  is 
allowed  to  enter  the  cylinder  only  during  a  part  of  the  suction 
stroke  at  light  loads.  The  length  of  time  the  mixture  is  ad- 
mitted is  controlled  by  the  governor.  With  the  throttling  gover- 
nor, the  normal  mixture  is  taken  in  during  the  whole  of  the  suc- 
tion stroke,  but  the  opening  is  throttled  so  that  not  as  much 
enters  at  light  loads  as  at  full  load. 

QUALITY  GOVERNOR.  This  governor  changes  the  ratio  of  the 
fuel  to  the  air  at  different  loads.  At  full  load,  a  rich  mixture  is 
used,  and  at  light  load  a  lean  mixture.  Mechanically,  this  scheme 
is  quite  simple,  but  it  has  the  disadvantage  of  giving  low  effi- 
ciency at  light  loads.  If  the  mixture  gets  too  rich  or  too  lean, 
it  may  be  impossible  to  secure  ignition.  Oil  engine  governors 
commonly  control  the  amount  of  oil  admitted  per  working 
stroke.  This  is  seen  to  be  the  same  as  the  quality-governing 
scheme. 

It  should  be  mentioned  that  the  speed  of  an  engine  can  be 
regulated  by  changing  the  time  of  ignition.  With  either  too 
early  or  too  late  ignition  the  full  power  is  not  developed  in  the 


GAS   ENGINES  211 

cylinder.  The  speed  of  motor-boat  engines  is  often  controlled 
in  this  way.  With  high  speeds,  the  spark  should  occur  earlier 
than  in  low-speed  engines.  With  a  variable-speed  engine,  such 
as  exists  in  an  automobile  or  truck,  the  time  of  ignition  should 
be  adjusted  to  the  speed  in  order  to  get  the  best  results. 

182.  Determination   of  Horsepower.  —  The  indicated   horse- 
power of  a  gas  engine  is  determined  in  the  same  way  as  for  the 
steam  engine,  with  the  exception  that  for  four-stroke  cycle  en- 
gines only  half  the  r.p.m.  is   used   in   the    computation.     If   a 
hit-and-miss  governor  is  used  on  the  engine,  the  number  of  hits 
per  minute  must  be  counted  rather  than  the  r.p.m.  of  the  shaft. 

183.  Multi-cylinder  Engines.  —  The  single  cylinder,  single-act- 
ing four-stroke-cycle  gas  engine  has  one  impulse  stroke  in  two  revo- 
lutions.    The  double-acting  steam  engine  has  two  impulse  strokes 
per  revolution.     Thus  it  is  seen  that  the  single  cylinder  gas  en- 
gine has  a  much  greater  variation  in  angular  acceleration  of  crank 
shaft  than  does  the  steam  engine.     For  some  kinds  of  service  this 
variation  in  angular  velocity  is  immaterial;  in  other  cases  it  is 
a  serious  disadvantage.     For  instance,  in  the  generation  of  elec- 
tric current  to  be  used  for  lighting,  a  single-cylinder  engine  is 
impracticable,  unless  an  exceedingly  heavy  flywheel  is  used.     To 
approximate  the  uniformity  of  torque  that  exists  in  a  single- 
cylinder  steam  engine,  it  is  necessary  to  use  four  cylinders  on  the 
gas  engine. 

In  automotive  service,  a  fairly  uniform  torque  is  desirable,  and 
therefore  four  or  more  cylinders  are  used.  If  more  than  four 
cylinders  are  used,  there  will  be  less  variation  in  the  angular 
torque,  and  the  engine  speed  may  be  controlled  more  easily  by 
the  throttle.  Too  large  a  number  of  cylinders  may  cause  a  de- 
crease in  the  efficiency  of  the  engine.  This  may  be  explained 
by  the  fact  that  with  a  multi-cylinder  engine,  there  is  more  area 
of  cylinder  wall  exposed  to  the  burned  gas  for  the  same  volume 
of  gas  than  there  is  in  the  single-cylinder  engine.  Principle  I  of 
§169,  as  set  forth  by  Beau  de  Rochas,  is  a  statement  of  this 
same  fact.  While  a  lowered  efficiency  may  result  from  the  use 
of  a  large  number  of  cylinders,  this  loss  may  be  more  than  com- 
pensated by  the  added  smoothness  of  running.  Many  of  the 
higher  grade  automobiles  made  at  the  present  time  are  equipped 
with  six,  eight,  or  even  twelve  cylinders. 


212 


ENGINES   AND   BOILERS 


PROBLEMS  213 

PROBLEMS 

1.  (a)  What  is  the  pressure  in  pounds  per  square  inch  that  corresponds  to 
a  mercury  column  16  inches  high? 

(6)  What  is  the  atmospheric  pressure  when  the  barometer  reads  27.4  inches? 

2.  A  steam  gage  is  used  to  show  the  pressure   in  a  steam  line  and  is  at- 
tached as  shown  in  Fig.  A.     If  the  small  pipe  leading  to  the  gage  is  full  of 
water  and  the  gage  reads  183  pounds,  what  is  the  pressure  in  the  steam  line? 

3.  If   the   pressure    gage   on   a 
boiler  reads   150  pounds  and  the 
barometer  reads  29.3  inched,  what 
is  the  absolute  pressure  in  the  boiler 
in  pounds  per  square  inch? 

4.  The  vacuum  gage  on  a  con- 
denser reads  27.2  inches  and  at  the 
same  time  the  barometer  reads  29.1 
inches. 

(a)  What  is  the  absolute  pres- 
sure in  the  condenser  in  pounds  per 
square  inch?     (6)  What  is  the  vac- 
uum-gage   reading    reduced    to   a  JTIG 
30-inch  basis? 

6.  A  condenser  with  its  air  pump  is  guaranteed  by  the  manufacturer  to 
produce  a  vacuum  of  28.5  inches  (on  the  basis  of  a  30-inch  barometer).  Dur- 
ing the  acceptance  test  the  barometer  read  28.73  inches.  What  should  be 
thevacuum-gage  reading  to  maintain  the  guaranteed  vacuum? 

Q)  A  boiler-feed  pump  is  located  14  feet  below  the  water  line  of  the  boiler. 
The  pump  draws  water  from  a  tank  located  7  feet  below  the  pump  cylinder. 
If  the  pressure  in  the  boiler  is  40  pounds  gage,  neglecting  friction  losses  due 
to  the  flow  of  water,  etc.,  what  is  the  least  total  head  the  pump  must  act 
against:  (a)  In  feet?  (6)  In  pounds  per  square  inch?  (c)  What  is  the 
least  height  of  water  level  in  a  standpipe  above  the  boiler  in  order  that  the 
water  will  flow  into  the  boiler  by  gravity? 

(Y)  Reduce:  (a)  A  temperature  reading  of  50°  Centigrade  to  the  corre- 
sponding Fahrenheit  reading.  (6)  A  temperature  of  320°  Fahrenheit  to 
Centigrade.  (c)  75  (great)  calories  to  B.t.u.  (d)  46  B.  t.  u.  to  calories. 
(e]  33000  foot-pounds  to  B.t.u. 

8.  If  1,576,000  B.t.u.  are  given  to  an  engine  in  an  hour  and  if  the  engine 
can  convert  6  per  cent  of  this  heat  into  work,  what  is  the  horsepower  of  the 
engine?     (One  horsepower  is  33,000  foot-pounds  per  minute.) 

9.  A  sample  of  Indiana  coal  gave  the  following  proximate  analysis: 

Moisture  =3.81%,  Fixed  carbon  =  76. 16%, 

Volatile  combustible  =  13.62%,          Ash  =  6.41%. 
The  same  sample  when  dried  gave  the  following  ultimate  analysis: 
Carbon      =84.26%,  Oxygen    =1.73%,  Sulphur  =  1.22%, 

Hydrogen=  4.38%,  Nitrogen  =  1.75%,  Ash         =6.66%. 

The  oxygen  calorimeter  gave  a  calorific  value  of  14682  B.t.u.  per  pound  of 
dry  fuel.  Find  the  calorific  value  of  the  preceding  sample  per  pound  of  dry 
fuel,  from  (a)  the  proximate  analysis,  (6)  the  ultimate  analysis. 


214  ENGINES   AND   BOILERS 


A  sample  of  West  Virginia  coal  gave  the  following  proximate  analysis : 
Moisture  =4.85%,  Fixed  carbon  =  68.36%, 

Volatile  combustible  =  16.31  %,          Ash  =  10.84%. 
The  same  sample  when  dried  gave  the  following  ultimate  analysis: 
Carbon      =80.34%,          Oxygen    =  3.11%,          Sulphur  =  .49%, 
Hydrogen  =  4.00%,          Nitrogen  =   1.05%          Ash          =11.01%. 
The  oxygen  calorimeter  gave  for  a  similar  dried  sample  a  calorific  value  of 
14,180  B.t.u.  per  pound.     Find  the  calorific  value  per  pound  of  dry  coal  from 
^  (a)  the  proximate  analysis  (b)  The  ultimate  analysis. 

(llj  Find  the  theoretical  weight  of  air  required  to  burn  completely  a 
pound  of  the  dry  coal  of  Problem  10. 

12.   A  boiler  test  was  run  using  the  coal  from  which  the  sample  of  Problem 
10  was  taken.     During  the  test  the  analysis  of  dry  flue-gas  was  as  follows: 
Carbon  dioxide  =  8.58%,  Carbon  monoxide  =     .05%, 

Oxygen  =11.32%,  Nitrogen  =  80.05%. 

Find  the  approximate  weight  of  air  used  to  burn  one  pound  of  dry  coal. 
(13?  In  the  test  of  Problem  12,  the  temperature  of  air  entering  the  furnace 
was'64°  F.,  and  the  stack  temperature  was  559°  F.     Find  the  percentage  of 
available  heat  carried  up  the  stack  by  the  dry  flue-gas. 

14.  Water  is  fed  to  a  boiler  at  a  temperature  of  170°  F.  The  pressure 
gage  reads  140  pounds,  and  the  barometer  29.6  inches.  How  many  B.t.u. 
are  needed  to  generate  a  pound  of  dry  steam?  What  is  the  temperature  of 
thesteam  generated?  The  volume  per  pound? 

Q$  Steam  at  a  gage  pressure  of  135  pounds  is  generated  from  water  at 
120°  F.  The  temperature  of  the  steam  is  490°  F.  The  barometer  reads 
29.3  inches.  Find  the  B.t.u.  required  to  generate  one  pound  of  steam. 

16.  If  the  temperature  of  steam  in  a  condenser  is  115°  F.,  what  is  the  great- 
estjaossible  vacuum-gage  reading,  if  the  barometer  reads  29.16  inches? 

Qj)  If  water  boiling  under  a  pressure  of  185  pounds  gage  is  allowed  to 
escape  to  the  atmosphere  (as  in  a  boiler  explosion),  what  percentage  of  its  weight 
turns  to  steam?  What  is  the  ratio  of  its  new  volume  to  the  old?  Assume 
that  the  barometer  reading  is  29.6  inches. 

18.  Dry  steam  leaves  a  boiler  at  a  pressure  of  180  pounds  gage  and  reaches 
the  engine  with  a  quality  of  98  per  cent,  and  a  pressure  of  177  pounds  gage. 
What  percentage  of  its  heat  contents  has  it  lost  in  its  passage  through  the 
pipe?  What  percentage  of  its  volume?  Assume  that  the  barometer  reading 
is  29.43  inches. 

Q9^  If  one  pound  of  steam  of  95  per  cent  quality  at  atmospheric  pressure  is 
mixed  with  8  pounds  of  water  at  70°  F.,  what  will  be  the  resultant  tempera- 
ture? Assume  that  the  barometer  reading  is  29.00  inches. 

20.  Dry  steam  enters  a  turbine  at  a  pressure  of  180  pounds  gage;  leaving 
the  turbine  it  passes  into  a  condenser  in  which  the  vacuum  is  27.6  inches  (30- 
inch  basis).  The  quality  of  steam  as  it  leaves  the  turbine  is  87%.  Neglect- 
ing all  losses,  find  how  many  foot-pounds  of  work  may  be  obtained  from  each 
pound  of  steam  that  passes  through  the  turbine. 

/2L)  A  frictionless  piston  weighing  7000  pounds  is  placed  in  a  vertical 
cyimcler  10  inches  in  diameter.  Two  pounds  of  water  at  70°  F.  are  placed 


PROBLEMS  215 

under  the  piston.     If  800  B.t.u.  are  added  to  the  water,  how  far  will  the  pis- 
ton move?     The  barometer  reads  29.6  inches. 

22.  If,  in  Problem  21,  2500  B.t.u.  are  added  to  the  water,  what  will  be 
the  weight  of  steam  formed?  What  will  be  its  temperature?  How  far  will 
thejMston  move? 

£2iP  A  sample  of  steam  is  taken  from  a  steam  line  in  which  the  pressure 
is  150  pounds  gage  and  is  led  to  a  throttling  calorimeter  in  which  the  tem- 
perature is  230°  F.  and  the  gage  pressure  is  3  pounds.  The  barometer  reads 
29.4  inches.  What  is  the  quality  of  steam  in  the  line? 

24.  A  horizontal  water-tube  boiler  (B.  and  W.  type)  has  10  vertical  rows 
of  four-inch  tubes,  9  tubes  to  the  row.  The  tubes  are  18  feet  long,  and  the 
steam  drum  is  24  feet  long  and  42  inches  in  diameter.  Find  the  heating 
surface  and  the  rated  horsepower  (a)  by  using  as  heating  surface  the  out- 
side surface  of  the  tubes  and  one-half  the  surface  of  the  drum;  (6)  by 
RuJle3,  p.  26. 

£j2J5c  A  horizontal  return  tubular  boiler  60  inches  in  diameter  and  18  feet 
long  has  44  four-inch  tubes.  Find  the  heating  surface  and  rated  horsepower 
(a)  by  Rule  1,  p.  26,  (6)  by  Rule  3,  p.  26. 

26.  A  Scotch  marine  boiler-shell  is  16  feet  3  inches  in  diameter  and  12 
feet  long.     There  are  three  furnaces,  each  43  inches  in  diameter.     The  boiler 
contains  three  sections  of  tubes,  each  section  consisting  of  110  three-inch 
tubes  10  feet  long.     Find  the  approximate  heating  surface  and  the  horse- 
power. 

27.  A  vertical  fire-tube  boiler  (exposed-tube  type)  has  a  diameter  of  30 
inches  and  a  height  of  6  feet.     The  furnace  is  25  inches  in  diameter  and  27 
inches  high.     There  are  55  two-inch  tubes  45  inches  long.     The  normal  water 
level  is  10  inches  from  the  top  of  the  tubes.     Find  the  heating  surface  and 
rated  horsepower  by  Rule  2,  p.  26. 

£^5P  In  a  test  of  a  B.  and  W.  boiler  with  a  hand-fired  furnace  at  the  Sew- 
age Pumping  Station,  Cleveland,  Ohio,  the  following  data  were  taken: 

Rated  horsepower  of  boiler 150 

Grate  surface 27  square  feet 

Duration  of  test 24  hours 

Steam  pressure 156.3  pounds  gage 

Temperature  of  feed  water 58°  F. 

Quality  of  steam  formed 99  per  cent. 

Total  weight  of  coal  fired  (wet) 15078  pounds. 

Moisture  in  coal 7.5  per  cent. 

Total  weight  of  water  fed  to  boiler 105100  pounds. 

Find: 

(a)  Factor  of  evaporation. 

(6)  Dry  coal  per  square  foot  of  grate  surface  per  hour. 

(c)  Equivalent  evaporation  per  hour  (from  and  at  212°  F.). 

(d)  Equivalent  evaporation  per  hour  per  square  foot  of  water-heating 
surface. 

(e)  Boiler  horsepower  developed. 

(£L  Percentage  of  rated  capacity  developed. 
(297  In  the  test  of  Problem  28,  the  dry  coal  had  a  calorific  value  of  12292 


216  ENGINES   AND   BOILERS 

B.t.u.  per  pound,  and  the  cost  delivered  at  the  boiler  room  was  $3.50  per 
ton  of  2000  pounds.     Find: 

(a)  Equivalent  evaporation  from  and  at  212°  per  pound  of  dry  coal. 
Combined  efficiency  of  boiler,  furnace  and  grate. 
Coal  cost  per  1000  pounds  of  equivalent  evaporation. 
In  a  test  of  a  B.  and  W.  boiler  the  following  data  were  taken: 

Rated  horsepower  of  boiler 508 

Grate  surface 90  square  feet 

Duration  of  test 16.25  hours 

Steam  pressure 199  pounds  gage 

Temperature  of  feedwater 48.4°  F. 

Superheat 136.5°  F. 

Total  weight  of  coal  fired  (wet) 39670  pounds 

Moisture  in  coal 4.22  per  cent 

Total  weight  of  water  fed  to  boiler 336200  pounds 

Find: 

(a)  Factor  of  evaporation. 

(6)  Dry  coal  per  square  foot  of  grate  surface  per  hour. 

(c)  Equivalent  evaporation  from  and  at  212°  per  hour. 

(d)  Equivalent  evaporation  from  and  at  212°  per  hour  per  square  foot 
of  water-heating  surface. 

(e)  Boiler  horsepower  developed. 
Percentage  of  rated  capacity  developed. 

The  coal  in  the  test  of  Problem  30  gave  the  following  proximate  anal- 
ysis when  dry:  volatile  combustible,  19  66  per  cent;  fixed  carbon,  75.41  per 
cent;  ash,  4.93  per  cent.  The  cost  delivered  to  the  boiler  room  was  S3. 75  per 
ton  of  2000  pounds.  Find: 

(a)  Equivalent  evaporation  per  pound  of  dry  coal. 
(6)  Combined  efficiency  of  boiler,  furnace  and  grate. 

Coal  cost  per  1000  pounds  of  equivalent  evaporation. 
Is  the  boiler  of  Problems  28  and  29  working  harder  than  that  of 
Problems  30  and  31,  or  conversely?     Give  the  reason  for  your  answer. 

33.  Find  the  size  of  a  pop  safety-valve  with  a  45°  seat  for  a  60-horsepower 
return-tubular  boiler  which  is  to  carry  a  gage  pressure  of  75  pounds.  Assume 
that  the  maximum  evaporation  is  5  pounds  of  water  per  hour  per  square  foot 
of  wetter-heating  surface,  and  that  the  lift  of  the  valve  is  1/30  of  the  diameter. 
(3p  How  many  2.5-inch  pop  safety-valves  would  one  4.5-inch  valve  re- 
place, assuming  that  the  lift  is  proportional  to  the  diameter? 

35.  How  many  3.5-inch  pop  safety-valves  are  required  for  the  boiler 
of  Problem  30?     Assume  the  rate  of  maximum  evaporation  as  6  pounds  of 
water  per  square  foot  of  water-heating  surface  per  hour,  and  that  the  lift 
is  1/30  of  the  diameter. 

36.  What  should  be  the   size  of  the  pop  safety-valve  for  the  boiler  of 
Problem  28 

(a)  Computed  as  in  Problem  35? 

(6)  Computed  from  the  P.  G.  Darling  formula?     See  p.  59. 

(c)  Computed  from  the  city  of  Chicago  formula?     See  p.  59. 

(d)  Computed  from  the  city  of  Philadelphia  formula?     See  p.  59. 


PROBLEMS  217 

(e)  Computed  from  the  U.  S.  Supervising  Inspectors'  formula?    See  p.  59. 

(/)    Computed  from  the  A.  S.  M.  E.  Boiler  Code  Committee's  require- 
menjt^?     See  Report  of  Boiler  Code  Committee  of  A.  S.  M.  E. 
/3T,/  What  should  be  the  size  of  a  steam  pipe  leading  from  a  250-horse- 
power  boiler  if  the  pressure  carried  is  160  pounds  gage?      Assume  a  velocity 
of  florin  the  pipe  of  5000  feet  per  minute. 

(ZSy/A  5000-kw.  steam  turbine  requires  16  pounds  of  dry  steam  per 
hour  per  kw.  at  160  pounds  gage  pressure.  The  vacuum  in  the  exhaust 
of  the  turbine  is  27.5  inches  of  mercury  (30-inch  barometer).  The  quality 
of  steam  in  the  exhaust  is  85%.  If  the  velocity  of  flow  of  steam  to  and 
away  from  the  turbine  is  to  be  7500  feet  per  minute,  what  should  be  the 
size _ol  steam  and  exhaust  pipes? 

/39.) If  a  steel  steam  pipe  is  to  carry  steam  at  a  pressure  of  200  pounds 
gageand  may  be  as  cold  as  30°  F.  when  the  steam  is  cut  off,  how  far  apart 
should  expansion  joints  be  placed  if  each  joint  gives  a  3-inch  movement? 

40.  If  9536  pounds  of  water  at  a  temperature  of  60°  F.  are  mixed  with 
1160  pounds  of  steam  at  3  pounds  gage  pressure,  the  steam  being  of  90  per 
cent  quality,  what  will  be  the  resultant  temperature  of  the  mixture? 
^41.  The  exhaust  from  a  65-horsepower  steam  engine  is  led  to  an  open 
feedwater  heater.  The  engine  uses  30  pounds  of  steam  per  hour  per  horse- 
power, and  the  quality  of  the  exhaust  steam  is  80%.  The  heater  is  at  atmos- 
pheric pressure;  water  enters  at  50°  F.  and  is  heated  to  200°  F. 

fa)  What  horsepower  of  boilers  will  the  heater  supply? 

(6)  What  should  be  the  size  of  steam  and  water  pipes  leading  to  the 
heater?  Assume  a  steam  velocity  of  5000  feet  per  minute  and  a  water  veloc- 
ity of  150  feet  per  minute. 

42.  A  4000-kw.    steam    turbine  is  equipped  with  a  surface  condenser 
The    turbine  uses  16  pounds  of  steam  per  kw.  per  hour,  which  enters  the 
condenser  at  a  quality  of  85  per  cent.     The  vacuum  to  be  maintained  is  28 
inches   (30-inch  basis).     The  circulating  water  enters  the  condenser  at  a 
temperature  of  60°  F.,  and  leaves  at  a  temperature  10°  cooler  than  that  of 
the  incoming  steam.      (a)  How  much  circulating  water  is  needed  per  hour? 

(6)  If  the  same  amount  of  water  is  circulated  as  in  part  (a),  but  if  it  enters 
at  90°  instead  of  60°,  and  leaves  at  10°  cooler  than  the  incoming  steam,  what 
vacuum  can  be  maintained? 

43.  An  18"X24"  steam  engine  has  a  piston  rod  2.75  inches  in  diameter. 
Find  the  head-end  and  the  crank-end  piston  displacements  in  cubic  feet. 

44.  If  it  takes  10.6  pounds  of  water  to  fill  the  head-end  clearance  space 
and  11.2  pounds  to  fill  the  crank-end  clearance  space  of  the  engine  in  Problem 
43,  what  is  the  percentage  of  clearance  for  each  end  of  the  engine? 

45.  Find  the  volume  of  steam  back  of  the  piston  of  the  engine  of  Problem 
43:  when  the  piston  is  at  12.4  per  cent  of  the  head-end  stroke;  when  it  is 
at  14.0  per  cent  of  the  crank-end  stroke. 

^  46.  Find  the  weight  of  dry  steam  back  of  the  piston  of  a  24"  X  36"  engine 
when  it  is  at  30  per  cent  of  the  head-end  stroke.  The  head-end  clearance  is 
4  per  cent  and  the  steam  pressure  back  of  the  piston  at  the  above  position 
is  105  pounds  gage.  If  we  know  that  at  this  time  there  is  actually  1.06  pounds 
of  wet  steam  back  of  the  piston,  what  must  be  its  quality? 


218 

* 


ENGINES   AND   BOILERS 


Construct  a  hypothetical  indicator  diagram,  using  the  following  data. 
Length  of  diagram  =4  inches  (this  does  not  include  clearance). 
Initial  pressure  =  150  pounds  per  square  inch  (gage). 
Back  pressure  =5  pounds  per  square  inch  (gag°). 
Cut-off  =  25  per  cent,  Release  =95  per  cent. 
Compression  =  15  per  cent,  Admission  =  2  per  cent. 
Clearance  =  7  per  cent. 

Atmospheric  pressure  =  15  pounds  per  square  inch, 
as  a  scale  of  pressure  60  pounds  per  inch. 

(Construct  a  hypothetical  indicator  diagram    for  a  uniflow   engine 
108),  using  the  following  data. 
Length  of  diagram  =4  inches.         Initial  pressure  =  170  pounds. 
Back  pressure  =  —  12  pounds  (engine  is  running  condensing). 
Cut-off  =  20  per  cent.  Release  and  compression  each  =90%. 

Admission  =  2  per  cent.  Clearance  =  3  p  r  cent. 

Also  show,  by  a  dotted  line  on  the  same  diagram,  the  compression  curve  when 
the  engine  runs  non-condensing  (back  pressure  =  0). 

State  in  what  ways  this  excessive  compression  may  be  relieved. 
49.   Compute  approximately  the  percentage  of  head-end  and  of  crank-end 

clearance  of  the  engine  from  which 
the  cards  of  Fig.  B  were  taken.  Use 
two  methods.  Cards  were  taken 
with  an  80-pound  spring. 

50.  Compute  the  engine  con- 
stant, or  the  horsepower  constant 
(LA/33000),  for  the  head  end  and 
for  the  crank  end  for  a  10"X14" 
engine  with  a  2"  piston  rod.  Your 
answer  must  be  correct  to  within 

onephalf  of  one  per  cent. 

-pIG  -g  (51/  Find  the  indicated    horse- 

power (i.  hp.)  of  the  steam  engine  of 

Problem  50  when  the  head-end  mean  effective  pressure  is  34.2,  and  the  crank- 
endgjke.p.  is  35.4  pounds  per  square  inch.  The  engine  is  running  at  260  r.p.m. 
(52/  A  test  was  run  on  a  14"  X 18"  steam  engine  with  a  2"  rod.  The 
head-end  m.  e.  p.  was  found  to  be  35.2  pounds  per  square  inch  and  the  crank- 
end  m.e.p.  34.6  pounds  per  square  inch.  The  speed  of  the  engine  was 
250  r.p.m.  The  power  was  absorbed  by  a  Prony  brake  whose  arm  is  6'  5" 
long.  The  effective  weight  of  the  brake  arm  on  the  scales  was  45  pounds. 
During  the  test  the  pressure  on  the  scales  was  382  pounds.  Find  (a)  the  indi- 
cate4\h°rseP°weiS  (&)  the  brake  horsepower;  (c)  the  mechanical  efficiency. 

H>3/  The  test  of  Problem  52  was  run  for  45  minutes,  during  which  time 
thVengine  used  2750  pounds  of  steam  at  a  pressure  of  120  pounds  gage,  and 
at  a  quality  of  97  per  cent.  Find: 

(a)  Dry  steam  used  per  indicated  horsepower  per  hour. 
(6)  B.t.u.  per  indicated  horsepower  per  minute, 

(c)  Thermal  efficiency  based  on  i.  hp. 

(d)  Thermal  efficiency  based  on  b.  hp. 


PROBLEMS 


219 


The  indicator  diagram  in  Fig.  C  and  the  following  data  were  taken  during 


cut-off 


re/ease 


compression 
FIG.  C 


a  test  of  a  Buckeye  engine. 

Size  of  engine,  7.75"  X 15",  \\\" 

rod. 
Radius  of  Prony  brake  arm  =6.02 

feet. 

Room  temperature  =  73.5°  F. 
Temperature  in  throttling  calo- 
rimeter =221.5°  F. 
Steam  pressure  at  throttle  =  128.7 

pounds  per  square  inch  gage. 
Steam  pressure  in  calorimeter  = 

1.125  pounds  gage. 
Barometer  =  28.5  inch. 
R.p.m.=  222.5. 
Net  brake  load  =  140  pounds. 
Scale  of  indicator  spring  =  80  pounds. 
Steam  used  per  hour  =  1161  pounds. 

54.    Find  the  m.e.p.  of  the  cards  by  the  mean-ordinate  method. 
65.    Find  the  indicated  horsepower,  head-end,  crank-end,  and  total. 
Find  the  brake  horsepower. 
Find: 

Mechanical  efficiency. 
Pounds  of  steam  per  i.  hp.  per  hour. 
B.t.u.  per  i.  hp.  per  minute. 
Thermal  efficiency  based  on  b.  hp. 

Determine  from  each  card  the  percentage  of  stroke  and  the  steam  pres- 
sure for  each  of  the  following  events : 
(a)  Cut-off. 
(6)  Release 
(c)   Compression. 

59.    Determine  the  weight  of  dry  steam  back  of  the  piston  for  each  end 

at  the  events  of  cut-off,  release,  and 
compression. 

60.  Find  the  amount  of  re-evapo- 
ration or  condensation  per  hour  dur- 
ing expansion. 

61.  Find  the  weight  of  dry  steam 
per  hour  per  indicated  horsepower 
accounted  for  by  the  cards. 


56. 
57. 

(a) 
(6) 
(c) 
(d) 
58. 


ZO  pound  s/orinq. 


FIG.  D 


62.  Combine  the  indicator  dia- 
grams shown  in  Fig.  D,  and  deter- 
mine the  diagram  factor.  The  cards 
of  Fig.  D  were  taken  from  an  8.02" 
X15"X24"  cross-compound  Corliss 


steam  engine,  running  at  85  r.p.m.  The  head-end  clearance  of  the  high-pres- 
sure cylinder  is  7.4  per  cent  and  the  head-end  clearance  of  the  low-pressure 
cylinder  is  6.01  per  cent. 


220 


ENGINES   AND   BOILERS 


Determine  the  size  of  cylinders  for  a  compound,  two-cylinder,  double- 
acting  steam  engine  (receiver  type),  assuming  the  following  data:  i.  hp.  =  120, 
r. p.m.  =  100,  cylinder  ratio  =  1/3,  piston  speed  =  600  feet  per  minute,  initial 
steam  pressure  =  140  pounds  absolute,  termin^ljgressure  of  hypothetical  dia- 
gram =14  pounds  absolute,  vacuum  =  24  inches  (30-inch  basis),  and  diagram 
f  actor  =  .85 

QJ4,/  In  a  certain  two-cylinder  compound  steam  engine  the  number  of  ex- 
pansions is  10,  the  initial  steam  pressure  is  120  pounds  absolute  and  the  back 
pressure  is  5  pounds  absolute.  The  receiver  pressure  is  30  pounds  abso- 
lute. The  cylinder  ratio  is  1  to  3.  Neglecting  clearance  and  piston  rods, 
compare  the  work  done  in  the  two  cylinders  and  the  stresses  on  the  two 
piston  rods. 

65.   Given  a  cross-compound  steam  engine,  show  by  means  of  a  graph  the 

l/j/ve  shown  in  mid-position 


f£- 

>- 

"4*                                          >l 

\  head 
end 

/ohfon 

£-*r 

cnwft  \ 

end 

FIG.  E 

variation  in  power  distribution  when  the  governor  varies  the  cut-off  equally 
in  each  cylinder  (choose  at  least  three  cut-offs). 

66.  Proceed  as  for  Problem  65,  but  assume  that  the  governor  varies  the 
point  of  cut-off  in  the  high  pressure  cylinder  only. 

67.  Consider  a  12"X18"  steam  engine   (section  of  cylinder  and  valve 
shown  in  Fig.  E),  with  the  following  given  data. 

Connecting  rods  6  feet  long. 

Val ve- travel  =  6  inches. 

Head-end  lead  =  crank-end  lead  =  .25  inch. 

Head-end  steam  lap  =  1.25  inches. 

Head-end  exhaust  lap  =  .5  inch. 

Width  of  port  =  1.75  inches. 

(a)  Draw  the  valve  on  its  seat,  the  crank  position,  the  eccentric  position, 
and  the  position  of  the  piston  in  the  cylinder  when  the  crank  is  on  head-end 
dead  center.     (Make  your  drawing  y±  actual  size.) 
(6)  Draw  the  same  parts  for  head-end  cut-off. 

(c)  Draw  the  same  parts  for  head-end  admission. 

(d)  Draw  the  same  parts  for  head-end  release. 

(e)  Draw  the  same  parts  for  head-end  compression. 

(/)   Determine  the  percentage  of  the  stroke  for  each  of  the  above  events. 


PROBLEMS 


221 


'Consider  a  14"X16 
L  aruTwith  the  following  data. 


engine,  running  over,  with  direct  slide-valve 


valve-travel  =  4    inches, 

lead  =  1A  inch,' 

steam  lap  =  1  inch,v 

exhaust  lap  =  K  inch/ 

/  (a)  Valve  ellipse.  Draw  the  crank  circle  K  actual  size,  and  about  the 
same  center  draw  the  eccentric  circle  full  size.  Choose  12  equidistant  crank 
positions  and  find  the  corresponding  eccentric  positions. 

For  any  crank  position  (as  C,  Fig.  F2),  the  piston  is  at  a  distance  x  from 
its  mid-position,  and  at  the  same  time  the  eccentric  is  at  a  distance  y  from  its 


Jfeam  /ap+/ead 


Crank  on 
head-end 
dead  center 


FIG.  F2 


FIG.  F4 


mid-position.  Plot  y  vertically  and  x  horizontally,  for  all  12  positions  of  the 
crank. 

Connect  the  points  thus  found  by  a  smooth  curve.  Label  on  this  diagram 
the  following  details:  the  crank  position  at  each  event  of  the  stroke,  the  lead, 
the  steam  lap,  the  exhaust  lap,  the  maximum  port-openings,  and  the  angle 
of  advance. 

(6)  Bilgram  diagram.  Draw  the  crank  and  eccentric  circles  and  choose 
12  equidistant  crank  positions  as  in  (a) .  For  each  crank  position  (as  C  in  Fig. 
F3),  draw  a  dotted  line  parallel  at  a  distance  y  from  the  crank.  The  inter- 
section of  these  dotted  lines  is  the  Bilgram  construction  point  P.  About  this 
point  P,  draw  in  the  steam-lap  and  exhaust-lap  circles. 

Show  on  this  diagram  the  crank  position  at  each  event,  the  lead,  the  steam 
lap,  the  exhaust  lap,  the  maximum  port-openings,  and  the  angle  of  advance. 

(c)  Zeuner  diagram.  Draw  the  crank  and  eccentric  circles  as  before,  and 
choose  12  equidistant  crank  positions.  Lay  off  radially  on  the  crank  from  the 
center  of  the  crank  circle  the  eccentric  displacement  y  (Fig.  F4) ;  connect  all 
points  thus  found  by  a  smooth  curve. 

Show  on  this  diagram  the  crank  position  at  each  event,  the  lead,  the  steam 
lap,  the  exhaust  lap,  the  maximum  port-openings,  and  the  angle  of  advance. 


222  ENGINES   AND    BOILERS 

69.  Consider  an  engine  with  the  following  given  data. 

Direct  slide-valve.  Head-end  steam  lap  =  1|  ". 

Engine  running  over.  Crank-end  steam  lap  =  1  inch. 

Valve-travel  =  5  inches.  Head-end  exhaust  lap  =  J4  inch. 

Head-end  lead  =  1/8  inch.  R/L  =  1/5. 

Find  the  head-end  and  crank-end  crank  positions,  and  the  percent  of  stroke 
at  each  event  by  means  of 

(a)  The  valve  ellipse,  (6)  the  Bilgram  diagram,  fc)   the  Zeuner  diagram. 

70.  Consider  an  engine  with  a  direct  slide-valve  and  with  the  following 
given  data: 

Engine  running  over.  Crank-end  cut-off  =50  per  cent. 

Valve-travel  =  3  inches.  Head-end  compression  =  25  per  cent. 

Head-end  admission  =  1  per  cent.    Crank-end  compression  =  25  per  cent. 
Head-end  cut-off  =  50  per  cent.      R/L  =  1/6. 

Find  the  percentage  of  stroke  at  all  events,  the  angle  of  advance  in  degrees, 
the  steam  laps,  the  exhaust  laps,  the  maximum  port-openings,  and  the  leads, 
by  means  of 

(a)  The  Bilgram  diagram,  (6)  the  Zeuner  diagram. 

Draw  the  eccentric  circle  full  size  and  the  crank  circle  to  such  a  scale  that 
it  is  the  same  size  as  the  eccentric  circle.  Label  all  of  the  dimensions  asked 
for  directly  on  the  diagrams,  also  label  the  head  end  of  the  diagram  and  in- 
dicate by  an  arrow  the  direction  of  rotation  of  the  crank. 

71.  Consider  an  engine  with  an  indirect  slide-valve  and  with  the  follow- 
ing given  data. 

Engine  running  over. 

Valve-travel  =  4  inches. 

Head-end  lead  =  |  inch. 

Crank-end  lead  =  Y±  inch. 

Head-end  cut-off  =  35  per  cent. 

Head-end  compression  =  15  per  cent. 

Sum  of  steam  lap  and  exhaust  lap  the  same  for  both  ends.     R/L  =  %. 
Find  the  percentage  of  stroke  at  all  events,  the  angle  of  advance  in  degrees, 
the  steam  laps,  the  exhaust  laps,  and  the  maximum  port-openings,  by  means  of 
(a)  The  Bilgram  diagram,  (6)  the  Zeuner  diagram. 

72.  Consider  an  engine  with  a  direct  slide-valve  and  with  the  following 
data. 

Engine  running  over. 

Head-end  admission  =  2  per  cent. 

Head-end  cut-off  =  60  per  cent. 

Head-end  maximum  port-opening  =  1  .25  inches. 

Crank-end  maximum  port-opening  =  1.25  inches. 

Head-end  compression  =  20  per  cent. 

Crank-end  compression  =  20  per  cent. 


Find  the  valve-travel,  the  angle  of  advance,  and  each  of  the  laps,  by  means 
of 

(a)  The  Bilgram  diagram,  (6)  the  Zeuner  diagram. 

Also  draw  to  scale  the  valve  on  the  seat  in  its  mid-position. 


PROBLEMS 


223 


73.   Consider  an  engine  with  a  direct  slide-valve  and  with  the  following 
data. 

Engine  running  under. 

Head-end  lead  =  J^  inch. 

Crank-end  lead=  f  inch. 

Head-end  cut-off  =  55  per  cent. 

Head-end  compression  =  20  per  cent. 

Crank-end  compression  =  20  per  cent. 

Head-end  maximum  port-opening  =  1.25  inch. 


Find  the  valve-travel,  the  angle  of  advance,  and  each  of  the  laps  by 
(a)  The  Bilgram  diagram,  (6)  the  Zeuner  diagram. 
Draw  the  valve  to  scale  in  mid-position. 


FIG.  G 

74.  Consider  an  engine  with  a  valve  whose  dimensions  and  seat  are  as 
shown  in  Fig.  G.  The  valve  is  not  shown  in  mid -position.  The  valve-travel 
is  4  inches;  R/L  =  l/5;  the  engine  runs  over. 

The  cards  of  Fig.  H  are  taken  with  the  valve  as  now  set.  Find  the  angle 
through  which  the  eccentric  must  be  shifted  (state  whether  backward  or  for- 


Cut-otf 


Cuf-off 


Crank-end  card 


FIG.  H 

ward),  and  the  amount  the  valve  stem  must  be  lengthened  or  shortened  (state 
which),  in  order  to  give  a  cut-off  of  25  per  cent  on  each  end.  Draw  the  ap- 
proximate cards  for  the  new  setting. 


224 


ENGINES   AND    BOILERS 


75.   Consider  an  automatic  shaft  governor  with  the  following  given  data 
(The  Rites  inertia  governor  is  shown  in  Figs.  I  and  J.) : 
Head-end  steam  lap  =  1.75  inches. 
Head-end  exhaust  lap  =  0. 

Lead  at  normal  position  of  eccentric  =  5/32  inch. 
Distance  of  eccentric  center  from  pivot  (R)  =  12". 
Distance  from  center  of  shaft  to  pivot  point  (x)  =  13f  ". 


FIG.  I 

Location  of  point  DE  =  25",  /=18". 
Location  of  point  A:  c  =  30",  6  =  52". 


I. 


Cut-off  at  no  load  =  10  per  cent. 
Cut-off  at  full  load  =  65  per  cent. 
Direct  slide-valve,  engine  running  over. 

Draw  the  governor  analysis  (full  size)  and  find  the  valve-travel  and 

angle  of  advance 

(a)  At  10%  cut-off,  (6)   at  65% 
cut-off,  (c)   at  normal  cut-off. 

(d}  Also  find  percent  of  normal 
cut-off. 

II.   Draw  the  head-end  Zeuner 
diagram,  or  the  Bilgram  diagram,  for 
(a)  10%  cut-off,  (6)  normal  cut- 
off, (c)   65%  cut-off. 

From  the  events  thus  deter- 
mined, draw  the  theoretical  indicator  diagrams,  using  6  per  cent  clearance, 
150  pounds  initial  steam  pressure,  5  pounds  back  pressure,  and  80  pounds  per 
inch  as  the  scale  of  spring. 

Find  the  elongation  of  governor  spring  (drawing  J/£  size) . 
(a)  From  no  load  to  normal,  (6)  from  normal  to  full  load. 


FlG 


PROBLEMS 


225 


76.  Consider  a  four-valve  engine,  such  as  is  shown  in  Figs.  60  and  61,  p. 
104,  with  head-end  valves  as  shown  in  Fig.  KI  and  K2,  and  with  the  follow- 
ing data. 

Radius  of  steam  valve  arms  =  5". 

Maximum  diameter  of  steam  or  cut-off  eccentric  circle  =  4". 

Diameter  of  exhaust  eccentric  circle  =  4". 

With  the  crank  on  head-end  dead  center,  the  pivot  point  of  the  governor 
arm  for  the  cut-off  eccentric  is  on  the  horizontal  center  line  8^"  beyond 
the  center  of  the  shaft.  Radius  of  locus  of  eccentric  centers  =  7  5/16". 

Cut-off  at  maximum  load  =  65  per  cent.       Compression  =  15  per  cent. 

Cut-off  at  normal  load  =  25  per  cent.  Release  =  95  per  cent 


va/ve  in  extreme  open  position 
(Fu//  /oad) 


'exhaust rtt/ye  !/r 
extreme  c/osed  joosift'on 

FIG.  K 

At  normal  load  the  steam-valve  arm  is  vertical  when  the  valve  is  in  extreme 
closed  position.  The  exhaust-valve  arm  is  vertical  when  that  valve  is  in  ex- 
treme open  position.  R/L  =  1/Q.  Engine  runs  over. 

Find  the  angle  of  advance  for  each  eccentric  at  normal  load. 

Find  the  location  of  the  steam-valve  arm  when  in  mid-position,  at  cut- 
off, at  admission,  and  when  it  is  in  extreme  open  position  at  normal  load. 

Find  the  maximum  port-opening  at  normal  load,  and  the  lead  at  normal 
load. 

Draw  the  head-end  steam  valve  in  extreme  position,  and  in  open  position 
at  normal  load. 

Draw  the  head-end  exhaust  valve  in  extreme  open  position. 


226 


ENGINES   AND   BOILERS 


77.   The  necessary  dimensions  of  a  Corliss  engine  are  given  in  Figs.  L 
and  M. 

Consider  such  an  engine  with  the  following  data. 

A  12"X24"  Corliss  engine  running  at  150  r.p.m. 

All  valves  operated  by  one  eccentric. 


Engine  runs  over. 


FIG.  L 

Diameter  of  all  valves =3"  (d,  Fig.  L). 

m,  Fig.  L=sy2"; 

k,  Fig.  L  =  15". 

Length  of  valve  arms  =  4'/. 

Center  of  wrist  plate  is  equidistant  from  all  valves. 

Radius  AO  and  BO  on  wrist  plate  =  5". 

Radius  HO  on  wrist  plate  (0  =6". 

Angles  EC' A'  and  FD"B"  =  90°. 

Release  =  98  per  cent. 


PROBLEMS  227 

Compression  =  4  per  cent. 

Crank  angle  at  admission  =  3°. 

Throw  of  eccentric  =  6f." 

Radii  of  rocker  arms  are  equal  for  eccentric  and  hook  rods. 

Normal  cut-off  =  20% . 

Width  of  admission  port  =  %". 

Width  of  exhaust  port  =  l". 

Single-ported  valves. 
In  Fig.  M, 

Radius  of  arm  EG  =3%' 

Radius  of  arm  EH  =4^". 

Radius  of  arm  El    =4". 

Radius  of  cam  EJ    =2". 

Radius  of  latch  EK  =  %1". 

Center  G  is  V/i"  above  horizontal  center  line  at  trip  position  for 

normal  cut-off. 

Make  the  general  layout  %  actual  size,  and  that  of  the  trip  mechanism 
full  size. 

Find   the   lengths   of   the   steam   rod  AC   and  the  exhaust  rod  BD,  the 
angle  of  advance,  the  steam  lap,  the  lead,  the  exhaust  lap,  the  maximum 


FIG.  M 

cut-off  with  trip  working,  the  maximum  cut-off  when  beyond  control  of  trip, 
the  maximum  port-opening  for  maximum  cut-off,  the  maximum  port-open- 
ing for  normal  cut-off,  the  maximum  port-opening  for  10%  cut-off,  the  move- 
ment of  the  governor  rod  from  normal  to  10%  cut-off,  and  the  movement  of 
governor  rod  from  normal  to  maximum  trip  cut-off. 


228 


ENGINES   AND   BOILERS 


PROBLEMS  229 

78.  The  necessary  dimensions  of  a  Stephenson  link  are  as  follows.   (Fig.  N.) 
Valve-travel  at  full  gear  =  5>£". 
Steam  lap  =  I". 
Exhaust  lap  =  Ty. 
Lead  at  full  gear=0". 
Steam  port  =  W. 
Exhaust  port  =  2Y2". 
Bridge  =  1". 
=  l/7.5 
a  =  b  =  45". 
m  =  3". 
e=U". 


h=U". 


.7  =  1713/32". 
k  =  l8". 
#==48.8". 
Make  the  drawing  one-half  actual  size  and  proceed  as  follows. 

(1)  Make  a  template  of  the  link  as  shown  in  Fig.  N. 

(2)  Locate  the  center  of  the  link-block  with  the  crank  at  head-end  dead 
center  at  full  gear. 

(3)  Find  the  center  of  travel  of  the  link-block,  neglecting  for  the  time 
being  the  angularity  of  the  eccentric  rods. 

(4)  Place  the  center  of  the  rocker  shaft  above  the  point  found. 

(5)  Place  the  crank  and  eccentrics  in  their  positions  at  40%  head-end  cut- 
off, running  forward. 

(6)  Find  by  trial  the  position  of  the  link  (template)  for  the  preceding 
position  of  the  crank.     (Remember  the  center  of  the  link-block  is  now  at  a 
distance  equal  to  the  steam-lap  from  its  mid-position.)     Locate  the  saddle- 
block  pin  and  the  position  of  the  bell  crank. 

(7)  Assume  twelve  equidistant  crank  positions  and  the  corresponding  ec- 
centric positions  for  the  preceding  cut-off.     Then  draw  in  the  center  of  link 
for  each  crank  position,  by  trial  by  means  of  the  template. 

(8)  Draw  a  valve  ellipse  from  the  valve  displacement  found  above.      Locate 
all  events. 

(9)  Check  as  to  the  assumed  cut-off. 

(10)  Find  the  amount  of  slip  between  the  link-block  and  link  when  running 
at  the  assumed  cut-off. 


230 


ENGINES   AND    BOILERS 


PROBLEMS  231 

79.  The  arrangement  of  parts  and  the  necessary  dimensions  of  a  Wal- 
schaert  gear  are  shown  in  Fig.  O.      The  valve  is  of  the  piston  type  and  has 
inside  admission.      This  is  the  common  type  used  on  modern  locomotives. 
Fig.  O  shows  the  piston  in  its  mid-position  and  the  link-block  set  at  mid-gear. 
As  the  link  is  pivoted  to  the  frame  of  the  engine  at  the  point  L,  there  will 
be  no  motion  of  the  link-block  when  set  at  mid-gear.     Hence  all  the  motion 
the  valve  gets  at  this  position  of  the  block  comes  from  the  cross-head.     There- 
fore, as  the  cross-head  is  in  mid-position,  the  valve  will  also  be  in  mid-position. 

Suppose  that  such  a  gear  is  used  on  an  engine  with  the  following  data. 

26J/TX30"  engine. 

Engine  runs  forward  (under). 

Diameter  of  valve  =  14." 

Maximum  valve-travel  =  6>£". 

Steam  lap  =  l  1/16". 

Exhaust  lap  =  0. 

Lead =3/16". 

Dimensions  as  in  Fig.  O. 
Make  the  drawing  one-fourth  actual  size,  and  proceed  as  follows: 

(1)  Lay  out  the  gear  in  the  position  shown  in  Fig.  O,  with  the  piston  in 
mid-position  and  the  link-block  set  at  mid-gear.     Indicate  each  of  the  parts 
by  its  center  line  only. 

(2)  Make  a  template  of  the  link. 

(3)  Set  the  valve  and  the  point  H  for  head-end  cut-off.     Then  place  the 
crank  at  40%  of  forward  stroke,  assuming  that  the  engine  is  running  forward. 

(4)  Locate  the  eccentric  E  at  90°  back  of  the  crank,  and  the  point  F,  and 
draw  in  the  center  line  of  the  link. 

(5)  With  the  cross-head  K  in  position  for  40%  cut-off,  the  location  of  the 
point  /  will  be  determined.     Since  H  was  located  in  (3),  the  point  G  is  found 
by  connecting  /  and  H.     The  distance  that  G  is  to  the  right  of  the  mid- 
position  gives  the  distance  that  the  link-block  center  is  to  the  right  of  its 
mid-position.    This  locates  the  link-block  for  40%  cut-off.    Now  locate  the 
points  D,  M  and  B. 

(6)  With  the  link-block  set  for  40%  head- end  cut-off,  take  twelve  equidis- 
tant crank  positions  and  find  the  valve  displacement  for  each  position.     Plot 
these   valve   displacements   against   the  corresponding  piston  displacements 
either  as  in  a  Zeuner  diagram  or  as  in  a  valve-ellipse  diagram,  and  connect 
the  points  thus  found  by  a  smooth  curve. 

(7)  Draw  in  the  laps  on  the  diagram  just  constructed,  and  check  the  cut-off 
with  the  assumed  value  of  40%. 

(8)  Find  the  amount  of  slip  between  the  link-block  and  the  link  when  the  posi- 
tion is  that  of  40%  cut-off,  with  the  engine  running  forward. 

80.  The  Russell  Engine  Co.  makes  a  four- valve  engine.     In  this  engine, 
the  exhaust  is  taken  care  of  by  oscillating  or  Corliss  valves,  and  the  admis- 
sion by  a  direct  slide-valve.     This  slide-valve  admits  steam  and  carries  on 
its  back  a  rider-valve  that  cuts  off  the  steam.    The  main  valve  is  driven 
by  an  eccentric  keyed  to  the  shaft,  while  the  rider-valve  is  driven  by  an 
eccentric  whose  angle  of  advance  is  controlled  by  the  governor.     The  gov- 
ernor simply  rotates  the  eccentric  about  the  shaft,  changing  the  angle  of 


232 


ENGINES   AND    BOILERS 


advance,  but  affecting  in  no  way  the  absolute  travel  of  the  valve.  Hence, 
it  is  necessary  to  consider  the  relative  motion  of  the  rider-valve  and  the  main 
valve  in  making  an  analysis  of  the  cut-off  valve. 

The  necessary  dimensions  of  the  valves  and  the  seat  are  given  in  Fig.  P. 
The  throw  of  both  eccentrics  is  5K  inches,  and  the  angle  of  advance  of  the 
main  eccentric  is  32.5  degrees. 

Proceed  with  the  analysis  in  the  following  manner. 

(1)  Draw  the  two  eccentric  circles  about  the  same  center  and   locate  the 
extremity  of  the  diameter  of  the  valve  circle  for  the  main  valve. 


Trgyf/  grf  both  IM/HCS  S4? 


Seaf 

FIG.  P    VALVES  AND  SEA^T  OF  RUSSELL  FOUR-VALVE  ENGINE 

(2)  Place  the  crank  for  cut-off  at  25%  of  the  head7end  stroke,  and  locate 
the  main  eccentric.     Then  determine  from  the  dimensions  in  Fig.  P  how  far 
from  the  mid-position  the  cut-off  eccentric  must  be  to  give  the  proper  posi- 
tion of  rider-valve  for  cut-off.     This  determines  the  angle  of  advance  of  the 
cut-off  eccentric  at  this  particular  cut-off. 

(3)  Take  twelve  equidistant  crank  positions  and  determine  the  relative 
position  of  the  rider-valve  to  the  main  valve  for  each  position.      Plot  these 
distances  as  in  a  Zeuner  diagram.     It  will  be  noticed  that  the  diameter  of 
the  relative  valve-circle  is  equal  in  amount  and  parallel  in  directio'n  to  a 
line  connecting  the  extremities  of  the  diameters  of  the  valve  circles  in  the 
Zeuner  diagrams  for  the  main  valve  and  the  rider- valve.      Also  determine 
the  relative  steam  lap,  which  is  negative.     It  will  be  found  that  this  is  equal 
to  the  distance  between  the  working  edges  of  the  rider-valve  and  the  main 
valve  when  both  are  in  mid-position. 


PROBLEMS 


233 


(4)  Now  determine  the  diameters  of  the  relative  valve-circles   in  amount 
and  in  direction  by  repeating  the  process  of  (3)  for  eight  cut-offs  (0%  to  70%), 
and  draw  the  locus  of  the  extremities  of  the  diameters  of  the  relative  valve- 
circles.     It  is  seen  that  this  locus  is  the  arc  of  a  circle  whose  radius  is  equal 
to  the  eccentricity  of  the  rider-valve  eccentric  and  whose  center  is  the  ex- 
tremity of  the  diameter  of  the  valve  circle  for  the  main  valve. 

(5)  Make  the  Zeuner  analysis  for  cut-offs  of  10%,  30%  and  60%,  finding 
the  relative  angle  of  advance  and  the  relative  valve- travel  by  drawing  a  per- 
pendicular to  the  crank  position  at  the  point  where  the  crank  cuts  the  rela- 
tive steam  lap.     The  point  where  this  perpendicular  intersects  the  locus  of 
the  extremities  of  the  diameters  of  the  relative  valve-circles  determines  the 
construction  point  for  the  relative  Zeuner  diagram. 

81.  A  gravity-balanced  spindle  governor  built  as  shown  in  Fig.  Q  has 
arms  20  inches  long.  At  normal  speed  the  arms  are  at  an  angle  of  45°  with 
the  horizontal. 

'  (a)  Find  the  normal  speed  of  the  governor. 

(6)  Find  the  percentage  of  variation  in  speed  from  no  load  to  full  load. 
(c)   If  the  normal  speed  is  increased  30  per  cent  and  the  range  of  vertical 
movement  of  the  point  A  is  the  same  as  before,  what  is  the  percentage  of  varia- 
tion in  speed  from  no  load  to  full  load? 

o 


-  S/o/oaJ 

-  formal 

-  -  FU///04J 


FIG.  Q 


FIG.  R 


82.  In  the  cross-armed  gravity-balanced  spindle  governor  shown  in  Fig.  R,. 
the  upper  arms  are  30  inches  long  and  the  lower  arms  are  20  inches  long. 

Find  the  percentage  of  variation  in  speed  from  no  load  to  full  load,  and 
compare  the  result  with  that  of  Problem  81. 

83.  The  governor  of  Problem  81  is  now  loaded  with  a  weight  of  60  pounds. 
If  the  normal  speed  is  now  100  r.p.m.,  and  the  vertical  movement  of  the  point  A 
is  the  same  as  before,  what  is  the  percentage  of  variation  of  speed  from  normal? 

84.  A  Corliss  engine  is  governed  by  a  loaded  gravity-balanced  spindle 
governor.     The  pulley  on  the  governor  and  pulley  on  the  engine  are  both 
10  inches  in  diameter.     It  is  desired  to  change  the  speed  of  the  engine  from 
100  to  120  r.p.m.      In  what  three  ways  may  this  be  done  without  affecting 
the  speed  regulation?    Give  your  calculations. 


234 


ENGINES   AND   BOILERS 


86.  Find  the  percentage  of  variation  in  speed  from  no  load  to  full  load  for 
the  governor  shown  in  Fig.  S. 

86.  Suppose  that  the  spring  of  a  spring-balanced  centrifugal  governor  is 
fastened  directly  to  a  weight  of  40  pounds,  as  shown  in  Fig.  T.  If  the  speed 


FIG.  S 


FIG.  T 


at  no  load  is  206  r.p.m.,  and  at  full  load  200  r.p.m.,  what  must  be  the  scale 
of  spring? 

87.  If  the  spring  in  Problem  86  is  replaced  by  one  of  40  pound  scale,  and 
the  speed  at  full  load  is  200  r.p.m.,  what  is  the  speed  at  no  load?     Is  the 
governor  stable  or  unstable? 

88.  What  scale  of  spring  would  make  the  governor  of  Problem  86  iso- 
chronous at  a  speed  of  200  r.p.m.?   • 

89.  The  no-load  speed  of  a  governor  is  300  r.p.m.  and  the  full  load  speed 

is  290  r.p.m.  At  no  load,  the  tension 
in  the  governor  spring  is  500  pounds. 
At  full  load,  the  spring  is  2  inches  shorter 
than  at  no  load.  The  scale  of  spring  is 
100  pounds. 

If  the  spring  is  tightened  by  shorten- 
ing it  up  half  an  inch,  what  will  be  the 
effect  on  the  no-load  and  full-load  speeds? 

90.  With  the  governor  of  Problem  89, 
how  much  would  the  spring  have  to  be 
taken  up  to  make  the  governor  isochro- 
nous?   What  would  the  speed  then  be? 

91.  In    the    steam-turbine    governor 
shown  in  Fig.  U,  the  weights  of  8  pounds 
each  are  4  inches  from  the  center  of  the 
spindle  at  full  load  when  the  speed  is  600 
r.p.m.  At  no  load,  the  weights  are  5  inches 


FIG.  U 


from  the  spindle  and  the  speed  is  610  r.p.m.  The  weights  are  pivoted  at 
points  A  and  A . 

(a)  Compute  the  scale  of  spring. 

(6)  Design  the  spring,  i.e.  find  the  size  of  spring  wire,  the  diameter  of  the 
coil,  and  the  number  of  turns  that  will  give  the  correct  force  and  scale. 


M512399        TJ255 
B9 
Forestry 


